~ , 

! 

I 

NASA -Cj(-/7~qqJt / 

DOE/NASAl0351-1 

NASA CR-174994 NASA-CR-174994 

19860015831 

Experimental Assessment 

of Advanced Stirling 

Component Concepts 

B. Ziph 

Stirling Thermal Motors, Inc. 

October 1985 

Prepared for 

MAY 1 ~: 1:1t56 

Ll\NGlEY RESEARCH CENTER 

LIBRARY, NASA 

HI'.,' '2TOl:l, YIRGINI~ 

NATIONAL AERONAUTICS AND SPACE ADMINISTRATION 

Lewis Research Center 

Under Contract DEN 3-351 

for 

U.S. DEPARTMENT OF ENERGY 

Conservation and Renewable Energy 

Office of Vehicle and Engine R&D 

, 111111111111111111111111111111111111111111111 

NF01229 
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publication, NTIS-PR-360 available from NTIS at the above address 
Experimental Assessment 

of Advanced Stirling 

Component Concepts 

B. Ziph 

Stirling Thermal Motors, Inc. 

Ann Arbor, Michigan 48104 

October 1985 

Prepared for 

DOE/NASAl0351-1 

NASA CR-174994 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

Under Contract DEN 3-351 

for 

U.S. DEPARTMENT OF ENERGY 

Conservation and Renewable Energy 

Office of Vehicle and Engine R&D 

Washington, D.C. 20545 

Under Interagency Agreement DE-AI01-77CS51040 
Section 

1 

2 

3 

4 

5 

TABLE OF CONTENTS 

Background and Introduction 

PI s ton Rmgs 

2.1 

2.2 

2.3 

2.4 

2.5 

Introduction and Summary 

SIngle, SOlId, Pressure-balanced PIston Ring 

Bonded PIston Ring Set 

Description of the Test RIg 

Energy Flow in the Test Rig 

2.6 Data ReductIon for PIston RIng Performance 

2.7 Results 

RecIprocatIng Rod Seal 

3.1 Introduction and Summary 

3.2 Rod Seal Design 

3.3 Descripti0n of the Tests 

Heat PIpes 

4.1 IntroductIon and Summary 

4.2 Heat Pipe ConstructIon and FIllIng Procedures 

4.3 DescrIptIon of the Tests and TheIr Results 

SplIt Shaft Component Design 

5.1 Introduction and Summary 

5.2 

5.3 

5.4 

Main Shaft Deflection Analysis 

BearIng LIfe CalculatIons 

Splme Strength 

AppendIX I 

Appendix II 

AppendIx II I 

PIston Ring Leakage Effects 

Piston Ring FrictIon 

Metallographlc Inspection of Heat PIpe #2 

111 

Page 

1 

6 

6 

9 

12 

18 

20 

27 

30 

46 

46 

48 

50 

56 

56 

57 

64 

70 

70 

74 

78 

82 

87 

93 

97 
SECTION I - BACKGROUND AND INTRODUCTION 

There are several StIrling engme technology concepts WhICh StIrlmg 

Thermal Motors, Inc. (STM) of Ann Arbor, MIchIgan, has advanced. These technologIes, If they achIeve theIr potentIal, should contnbute sIgmflcantly to the 

Department of Energy (DOE) funded AutomotIve Stirhng engme (ASE) program. 

Some of the most promIsmg technologIes WhICh STM has proposed for mvestigatIOn mclude: 

1. Press un zed crankcase wIth rotatmg shaft seal 

2. Advanced pIston rmgs WIth Improved performance 

3. PIston rod scraper seal 

4. Heat pIpes to transport energy into the engme 

The current ASE relIes on reciprocatmg pIston rod seals to contam the 

workmg gas. The hfe potentIal of these seals has not been demonstrated to 

meet the 3500 hour goal of the ASE program. STM has proposed an alternate 

approach to dynamIc seals currently not bemg pursued m the ASE program, where 

the crankcase IS pressurized to the mean cycle pressure and the workmg gas IS 

con tamed by a hIgh pressure rotatmg seal on the engme output shaft. These 

rotatmg seals are commercIally avaIlable and have a manufacturer's claImed 

hfe potential of more than 10,000 hours. For thIS concept, a reciprocatmg 

pIston rod seal consIstmg of a cap seal and 011 scraper IS stIll needed. But, 

It IS needed only to Isolate the varymg cycle pressures from the constant 

crankcase pressure and to prevent 011 from entering the thermodynamIc cycles. 

The use of a pressurized crankcase, however, precludes controlhng engme 

power by varymg the engme pressure as currently IS done m the ASE - mamly 

because of the comphcatIOns of handhng the 011 entramed as a mIst m the 

workmg gas m the crankcase. Therefore, an alternate means of power control IS 

1 
used where the pressure remams constant and the pIston stroke IS vaned. The 

vanable stroke can be accomplished by usmg a vanable angle swash plate dnve. 

STM has developed a desIgn for the system needed to control the vanable stroke 

dnve whIch IS qUite sImple and has a potentIal cost and reliabIlity advantage 

compared to the system requIred to vary engme pressure. The vanable swashplate drive also has the additiOnal potentIal advantages of bemg self-startmg 

and of havmg better part load engme effIcIency than a fIxed stroke engme. 

The mherent charactenstics of the swashplate dnve are such that the 

mam dnve shaft IS subjected to large bendmg moments. As a result the unsupported end of the shaft at the rotatmg seal locatIon tends to deflect more 

than the amount the seal can tolerate wIthout affectmg life and leakage rate. 

Therefore, a means IS needed to elimmate shaft run out at the rotatmg seal. 

STM has proposed such a desIgn usmg a split shaft. 

Another area of mterest m the ASE program IS pIston rmgs. PIston rmg 

fnction life and sealing efficmecy are Important factors m meetmg the ASE 

fuel economy goals. STM has proposed the use of a solid smgle pIston rmg (the 

current ASE uses two rmgs) wIth a means for controlling the contact force 

against the cylinder wall whIch potentIally would produce an Improvement m 

engine fuel economy. STM has also proposed a number of alternatIve piston-rmg 

desIgns and concepts all aImed at brmgmg about Improvement m engme effICIency through reduced fnctiOn and leakage m the pIston rmgs. 

The use of heat pIpes offers a number of potential advantages for the ASE 

as well as for other Stirlmg engme applicatIons: 

• It provIdes consIderable flexiblity to allow for packagmg reqUIrements and ease of manufacturmg. 

• It allows major SImplifIcatiOn of heat exchanger desIgn. An engme 

heater, desIgned to take advantage of the hIgh film heat transfer 

2 
coefficient of condensmg metal vapor can be considerably less expenSive and more sUltable for mass productIOn than conventIOnal heaters 

such as those used for the ASE. 

• It bnngs about considerable improvement to the engme performance by 

permittmg the heater design to be ideally sUited to the thermodynamic 

reqUirements. 

• The uniform temperature throughtout the conflOes of the heat-pipe 

enclosure elimmates hot spots on the heater and thus enhances both 

the effiCiency and the reliability of the engme. 

ConventIOnal heat pipe designs reqUire a lengthy procedure for chargmg. 

The procedure reqUires evacuatIOn to a hard vacuum and bake out at elevated 

temperature to remove gases entrapped withm the surface of the metal contamer. 

ThiS process generally is not sUlteable for volume productIOn as required for 

an ASE. The use of a getter withm the heat pipe, as proposed by STM, may 

elimmate or reduce the bake out reqUirement and elimmate the need for speCial 

productIOn equipment. Also, the use of a mechamcal couplmg to Jom heat pipe 

sections also proposed by STM would allow the sections to be made as Simple 

subassemblies, thus greatly reduclOg the potential part rejectIOn rate. 

In an unsoliCited proposal, Stirling Thermal Motors, Inc. has proposed to 

do analysis, deSign, fabncatIOn and test 109 of components meetmg the needs of 

the advanced concepts discussed above. Under the resultmg Contract No. DEN3-351 

the followmg areas were mvestigated: 

1) Piston Rmgs 

Two advanced piston rmg concepts were expenmentally assessed and 

compared to conventIOnal piston rlOgs (the performance of which was 

3 
measured on the same rig to establIsh a base-line). 

• Single, solid pIston nng WIth a bI-metal spring proved 

ImpractIcal In VIew of Its very hIgh sensItIvIty to the 

operating temperature. 

• A bonded pIston nng set, In prelIminary testing, proved a very 

prOmISing concept exhIbIting relatIvely low losses. 

2) RecIprocating Rod Seal 

A babbit scraper mounted In a complIant housing was subjected to long 

term endurance tests at atmosphenc pressure. Test condItIons provIded lateral motIOn of the rod whIch IS of about tWIce the amplItude 

expected In an engine. The seal operated for 6750 essentIally 

trouble-free hours. 

Use of the scraper seal as a gas pressure containment seal was 

also investIgated expenmentally. Whereas It dId not perform 

adequately as a containment seal the tests pOinted out some 

modifIcatIOns WhICh may adapt It successfully to such applIcatIon. 

3) Heat PIpes 

Twelve heat pIpes were buIlt and SIX were subjected to long-term 

endurance testing, clearly indIcating the adequacy of the sImple, 

ineXpenSIve fabncation and fIlling procedures used. In addItIon the 

tests proved the effectIveness of getters In rendering the heat pIpe 

rather tolerant to aIr leakage and making It possIble to avoId 

furnace degassmg dunng fabncatIOn. 

CommerCially avaIlable mechamcal couplIngs were demonstrated to 

be applIcable to the heat pIpes. They proved tolerant of thermal 

4 
cycling and able to readIly provIde capIllary connectIon. 

4) SplIt Shaft Component DesIgn 

A splIt shaft wIth a separate end seal assembly was desIgned In 

detaIl based upon a reference desIgn (STM4-120 variable swashplate 

StIrling engine) prevIOusly generated by STM. The desIgn fully meets 

all the criteria, solves the shaft run out problem at the seal area 

and ensures component lIfe greater than 10,000 hours at full load. 

The advanced concepts tested, the tests themselves, and theIr results are 

described In detaIl In thIS report. The follOWing SectIons 2 through 5 describe 

the pIston rings tests, rod seal tests, heat pIpe tests, and splIt shaft 

desIgn, respectIvely. 
SECTION 2 -PISTON RINGS 

2.1 - IntroductIOn and Summary 

Whereas the tradItIOnal arrangement of two UnidIrectIOnal (lIpumpIngll) 

pIston rings shown In FIgure 2.1 works well and reliably It does gIve rise to 

relatIvely hIgh frictIOn and leakage losses. These losses have been generally 

accepted as the unaVOIdable penalty WhICh one has to pay In order to have hIgh 

reliabIlity, long life and the VIrtual eliminatIOn of net mass transfer across 

the pIston rings whIch characterise the UnidIrectIOnal ring arrangement. 

The program described below, challenging the ineVItabIlity of the frictIOn 

and lea\<age losses, was aImed at the development of an alternatIve ring 

arrangement WhIch wIll reduce these losses. 

At the beginning of the program It appeared that the concept of a Single, 

solid, pressure-balanced pIston ring backed Up by a bImetallIc spring showed 

great promIse of reducing both frictIOn and leakage whIle effectIvely eliminating net mass transfer. ThIS ring, descrIbed In SectIon 2.2 below was tested 

WIth rather dISappOinting results. 

It turned out that the stIffness of the closed ring played a major role, 

and, In cOmbinatIOn WIth the large difference between the coeffICIent of thermal expanSIOn of the ring material (Rulon LD) and that of the steel cylinder, 

YIelded totally unreproducible results. The performance was so senSItIve to 

temperature and fabricatIOn tolerances (of the OD) that after testing with 

vanous modifIcatIOns It became eVIdent that a solid ring cannot be used adventageously or, at least, that some baSIC new approach to such a ring IS reqUIred. 

Since such baSIC new approach remained elUSIve the effort was concentrated 

on modIficatIOns to the split UnidIrectIOnal rings arrangement WIth the purpose 

6 
To plenum , 

PIston rmg WIth step Jomt 

PIston 

,-- PIston rmg 

/ 

Back-up rmg 

--Sprmg 

--rFigure 2.l 

TradItIOnal arrangement of two UnidIrectIOnal "pumpmg"piston rmgs (schematIc) 

7 
of reducing frictIOn and leakage. 

It was postulated that the splIt rIngs allow leakage predomInantly through 

the step JOint. An effort focused on the JOint, If successful In reducing the 

leakage thereat, wIll make it possible to reduce the aXial thickness of the 

ring without adversely effecting the leakage, and thus wIll reduce frictIOn. 

A test program was, consequently, defined aimed at proving the following: 

1) Leakage losses are governed by the step-Joint qualIty and are, by and 

large, Independent of the aXial thickness of the ring. 

2) FrIcton losses are proportional to the aXial thickness of the 

rings. 

In order to prove the above statements, frictIOn and leakage data were 

measured for the following three sets of rings: 

1) Ring Set II I was fabricated by Shamban to STM dimensIOnal 

specifications USing common fabricatIOn procedures. No particular 

attentIOn was paid to the flatness or smoothness of the step JOint. 

These rings had an effective aXial thickness of 1.5 mm. 

2) Ring Set 112 was made In-house. The step-Joint surfaces were 

carefully cut and polIshed. The aXial thickness was 2.4 mm. 

3) Ring Set 113 IS the same as RIng Set 112 except that the aXial 

thickness IS 3.7 mm. 

The results clearly demonstrated that the qualIty of the step JOint surface had a major Influence on the leakage, that the aXial thickness of the 

rings had only a mInor Influence on the leakage and that, as expected, the 

frictIOn IS proportIOnal to the aXial thickness of the ring. 

In accordance with the above conclUSions, a fourth set of bonded rings, 

described In detaIl In Section 2.3 below, was made and tested. These rings 

8 
are aXIally thin (I.7mm) and theIr step JOints are well sealed. 

PrelIminary results demonstrate that, as expected, the frictIon IS low, In 

accordance wIth the small aXIal thIckness, and the leakage IS not hIgher than 

that of much thIcker rings. These rings (hereinafter "bonded rings") show 

sIgnifIcant promIse. In additlOn to theIr hIgh performance potentIal they 

reqUIre only a Single groove In the one-pIece pIston and can accomodate a large 

amount of wear WIth no adverse Influence on theIr performance. It IS strongly 

recommended that further development of the bonded rings be undertaken. 

2.2 - Single, SolId, Pressure-Balanced PIston Ring 

The Single, solId, pressure-balanced pIston ring IS shown schematIcally In 

FIgure 2.2. Its deSIgn addresses the follOWing Issues: 

a) The ring being closed, and thus haVing no JOint, It should reduce the 

leakage substantIally since, as was stIpulated and later proven, It 

IS the Jomt that provIded the most Important leak path. 

b) The aXIal face of the ring being uninterrupted, smce the ring IS 

closed, can be POSItIvely sealed WIth an O-ring, elIminating the 

mechanism that may cause net mass transfer across the ring. Since 

the ring WIll not permIt net mass transfer only a Single ring IS 

necessary rather than a set of two. 

c) The geometry of the rmg ensures that, practIcally, no net force IS 

exerted on It by the gas pressure as shown In FIgure 2.3. The 

contact force at the cylmder wall may be prOVIded by a back-up 

spring. That force could be rather small and hence the fnctlOn 

loss WIll be mmInal. Furthermore, that low fnction IS mdependent 

~f the aXIal thIckness of the rmg WhICh may thus be rather large to 

reduce leakage even further. 

9 
Closed, L-shaped---+ 

PIston ring 

Cylinder wall 

.' 

.".--- O-ring 

,/ BI-metal spring 

-;;--.?' ~/ J-==~~~~-------~~~~' 

\ 

,\ \ 

\ , \\ \ \ \ \ \ \ \ \ \ 

\\,,\ '\ , \ \ \ \ \ \ \ \ '\ , \ \ \ .. e_, __ 

, \ \ " \ 

'. \ \ \ \ \ 

\ '\' 

. \ \ \ 

. \ \ \ 

.----.,.- PIston 

Fip;ure 2.2 

Single, solId', pressure balanced pIston ring wIth bI-metallIc back-up spnng 

Pabove Pabove 

f 

h 

I 

Pbelow h 4-+ 

04- h ~ 

~ i 

Fieure 2.3 

Balancing of pressure load on the balanced pIston ring 

10 
Any closed pIston ring IS necessarily assocIated wIth a cold-start faIlure 

unless some actIon IS taken to prevent such faIlure. The problem IS caused by 

the coeffIcIent of thermal expanSIon of the ring material WhICh IS much larger 

than that of the cylInder. As the operating temperature Increases, the cylinder wall restrainS the pIston ring and prevents It from expanding. The ring 

develops compreSSIve stress WhIch after some tIme at the operating temperature 

wIll be relIeved. Upon shut-down and coolIng of the ring to ambIent temperature the ring WIll shnnk diametncally and lose contact WIth the wall. 

SealIng could not be estabhshed upon subsequent start-up attempt and the 

engine would not start. 

To prevent thIS cold-start faIlure the pIston ring IS backed up by a bImetallIc spring as shown In FIgure 2.2. Upon cold-start, at low temperature, 

the spring force WIll force the pIston ring against the wall, ensuring successful start-up. As the operating temperature Increases and the pIston ring 

expands the free dIameter of the bI-metal spring wIll contract resulting In a 

reductlOn of the contact force so that at operating temperature only a small 

force WIll remain. The fnctlOn under operating condItIons wIll be minImal. 

Testing of the Single, sohd, pressure-balanced pIston rings WIth the bImetal spnng revealed that, In practIce, the cOmbInatlOn of the stIffness of 

the ring Itself, the bI-metal spring constant and the tolerence of the dimenSlOn of the outer dIameter of the pIston ring created a situatlOn where the 

behavlOr of the ring IS extremely senSItIve to the operating temperature. 

In some cases as the operating temperature was nSIng from start-up compreSSIve stress In the ring and the attandant nse In fnction drove the 

temperature stIll hIgher In an unstable loop WhICh resulted In the fnctlOn and 

temperature continually increasing at constant operating speed and pressure. 

In other cases the clearance gap between the ring and the cylinder wall 

11 
was too large and sealIng could not be establIshed. Consequently, this concept was abandoned as Impractical. 

2.3 - Bonded Piston RIng Set 

FunctlOnally, the bonded rIngs are Identical to the traditional splIt 

"pumpIng" piston rmgs which are employed as sets of two with maximum pressure 

In between. 

The specIfiC design of the compound rIngs addresses those Issues derIved 

from the very small aXial thickness of these rIngs. In particular, achievIng 

a good quahty step JOInt ensurIng that the low rIgidity of the rIngs IS not 

harmful to their proper operatlOn, and good sealIng of the back Side of the 

rIng are some of the maIn Issures addressed by the design. 

Other Important features are the abilIty of the rIng to undergo conslderable wear with no adverse effect on the performance and Simple, Inexpensive 

constructlOn of the rIng and the piston. 

The rIng set assembly IS shown schematically In Figure 2.4. The entire 

assembly fits Into a sIngle groove In the sohd (one piece) piston. The two 

rIngs are separated by a spacer rIng made of plastiC which Incorporates, on 

both ItS aXial faces, radial leakage grooves which prevent these faces from 

I 

sealIng when the Instantaneous cycle pressure IS higher than the pressure 

between the rmgs. The spacer rmg IS open to allow mstallatlOn In the 

groove of the one-piece piston. 

At the back of each rIng IS a back-up rubber rIng which prOVides sealIng 

of the radial leak path at the step JOInt. The edges of the step JOInt are 

not radial lInes, but are contoured so that at the outer dIameter the edge IS 

radial and at the Inner diameter It IS substantially tangential. ThiS Will 

prevent the rubber back-up rIng from extrudIng Into the slot after It has 

12 
Spring 

Back-up rubber ring 

Spacer 

Bonded pIston ring 

FIgure 2.4 

Bonded PIston Rmg Set 

13 
opened due to wear as shown In FIgure 2.5. ThIS WIll permIt the pIston ring 

to accomodate a substantIal amount of wear wIth no adverse effect on the 

performance. 

Rulon tape IS glued to the cylinder dIrectly above and below the pIston 

ring groove to prevent metal-to-metal contact between the pIston and the cylinder as well as to prevent the thin pIston rings from being extruded Into the 

gap between the pIston and the cylinder wall. Slanted leakage grooves In the 

Rulon tape, as shown In FIgure 2.6, prevent It from partIcIpating In the 

sealing process. A hIgh quality step JOint IS ensured by the method of fabricatIon of the bonded pIston rings as shown In FIgure 2.7. 

To fabricate each bonded ring two IdentIcal closed rings are made from 

thin Rulon sheet and split wIth a contoured tool (step 1 In FIgure 2.7). Glue 

IS applied to one ring covering ItS aXIal face except for the part Intended for 

the step JOint whIch IS left exposed. The other ring IS glued to the first 

ring so that the splits are not alligned but, rather, span the step JOint area 

(step 2 In FIgure 2.7) YIelding the finIShed compound ring. In productIon the 

rings wIth the contoured splIt could eaSIly be stamped from thm Rulon sheets 

or cast In one pIece. 

The entIre ring assembly, viz. the pIston rings, spacer ring, back-up 

rings and springs, can be mstalled m a single groove In the solId, one-pIece 

pIston. 

The space at the back of the rmg assembly IS connected to a convenient 

plenum to mInimIZe the pressure fluctuatIOn therein. A reasonably constant 

pressure (equal to the maXImum cycle presure) In that space IS essentIal for 

the proper operatIon of the ring set. 

The confIguration of the set of bonded pIston rings whIch were tested 

dIffered from the one described above In that the edges of the step Jomt were 

14 
.... --- New pIston rmg 

1 1 

Contoured edge StraIght radIal edge 

I 

.4----Worn pIston rmg __ -.. 

Figure 2.5 Advantages of Contoured Edge of Step Joint 

After the pIston rmg has worn and the step Jomt opened up, the 

contoured edge of the step Jomt prevents the rubber back-up 

rmg fron extrudmg mto the opemng. 

15 

1 
Figure 2.6 

Slanted leakage grooves m Rulon tape bearmg 

16 
,...... 

-...J 

Step 1 

Two rings with contoured slots 

cut from thm sheet 

Step 2 

Adhesive (shaded areas) appl1ed 

and rmgs are bonded with slots 

spannmg the step Jomt area 

Figure 2.7 

A fmished bonded~on ring :....:......::..:..:...:..;....;..-----

Method of fabrication of a bonded piston rmg 
radIal rather than contoured and the grooved spacer rmg was made of brass 

rather than plastIC. A two-pIece pIston was used. 

The pIston was equIpped wIth a dome, the mternal volume of whIch was 

connected to the back of the pIston rmg groove provIdmg a plenum of 38 cm ~ 

In testmg, the bonded pIston rmgs exhIbIted low frictIOn accordmg to 

theIr small aXIal thIckness. At the same tIme, the quahty of the step Jomt 

and the posItIve seahng of itS back side mamtamed relatively low leakage. 

The bonded rmgs, m prehmmary testmg, were demonstrated to be a very 

promIsmg concept whIch should be developed further. 

2.4 - DescriptIOn of the Test Rig 

The test rig IS shown schematIcally m FIgure 2.8. A double actmg, 

crank driven pIston IS compressmg the hehum gas m a sealed, blmd cyhnder. 

The variable volume above the pIston IS approXImately equal to that below the 

pIston. PIston rmgs mounted m the pIston separate the two variable volumes, 

whIch undergo roughly harmOniC pressure VariatIOn 180 degrees out of phase from 

each other. 

The part of the cyhnder whIch IS swept by the pIston ring IS equipped 

wIth a water Jacket. The water flow through the Jacket IS measured SImply wIth 

a stop watch and graduated cyhnder. The temperature difference between the 

mlet and outlet of the Jacket is measured with a thermoplle. In thIS way the 

energy mput to the water Jacket IS mdicated. 

Thermocouples m the cyhnder wall mdicate the wall temperature of the 

two variable volumes. 

Each of the variable volumes is equipped wIth two check valves, mounted m 

OppOSIte dIrectIOns, and wIth a restrictIOn. These are connected to pressure 

18 
f--' 

<D 

11 

12 

6 6 

13 

0-

10 

15 16 

Figure 2.8 

Variable Speed, Reciprocating Test Rig, Schematic 

1) Crankshaft 7) Water mlet 13) Mmlmum pressure gage 

2) Cross-head 8) Water outlet 14) Couphng 

3) Cylmder 9) Thermopile 15) Variable speed DC motor 

4) Rod seal 10) Water flow meter 16) Speed mdlcator 

5) Water Jacket 11) MaXimum pressure gage 17) Torque arm 

6) Thermocouples 12) Mean pressure gage 18) Balance 
gages WhICh indIcate the maXImum, minImUm and mean pressure In each of the two 

volumes. 

The rig IS driven by a variable speed DC motor mounted on bearings such 

that ItS houSing IS free to rotate as shown In FIgure 2.8. A torque arm 

attached to the housing IS equIpped wIth a knife edge WhICh presses on the tray 

of a balance. The balance reading IS proportIOnal to the driving torque, the 

proportIOnahty constant being the length of the torque arm. The speed IS 

measured by a magnetIc proXImIty sensor and cog wheel arrangement. The temperature of the lubricating 011 supphed to the crank drive of the rig IS 

maintained constant at 45°C by a thermostat controlled heated sump so that the 

effect of temperature on the drive frictIOn through ItS effect on the 011 

VISCOSIty IS ehmInated. 

A layout drawing of the rig IS shown In FIgure 2.9. FIgure 2.10 IS a 

photograph of the rig. 

In summary, the rig IS Instrumented to measure the following quantItIes: 

• MaXImum, minImUm and mean pressure In each of the two variable 

volumes; 

• RotatIOnal speed of the rig; 

• Mechanical energy Input to the rig; 

• Heat energy Input to the water Jacket; and 

• Wall temperature of each variable volume 

2.5 - Energy Flow In the Test RIg 

At steady state, mechanical energy (W) per revolutIon supphed to the rig 

and measured by the balance, IS converted In ItS entIrety to heat. Four 

separate mechanisms facIhtate thIS process: 

20 
N 

t--' 

3)- 111-11 

I') 

Figure 2.9 

RECIPROCATING TEST RIG 

24- 2 

H 2 

I 2. 8 

12/ e 1'0 Z 

" 8 18 -4 

17 ,[" Iv 2 ," 4 

,4 

I~ " '2 

,'1 '1 

" .! 

.0 '2 

~ .. f> '2 , '1 

U> 1-

.. '2 

.. ,s 

~ 2 

2 Z 

/ of 

D'" ""T 

O',toJ& 2-/4 PUR 

o-RJ~ 1.-14f fiJi? 

""'02 MIT 'TO 

... ,/ NUT 5TD 

P/Io/ "-0010 IOU> 

H!O'/ ""'0 35 U; ,"-0 

lit"" '01:.5 8 L6 4·" Src, -"",,2.'." wr sro PIN <0000' 10<0 

() ~'NCi 2..-215 _VIC 

MtJ"2~ 5TU)~ 31'",,, 'T. 

A55r"CoOOOa 

.55~ "'-0007 

P/>/ ,;;ooD{' .~U'A'" 

u " .~ P"" (...Q ...... a 

""::::... .;:;:~ .. "'''' "va 

p"", '0004' vA'" 400'" ",.0 Q,"", Roo-~~ .... ;0..,"-

.... ~.., .. ,OO'!>~ 

'" N «JOoS NlTRAUtJf6 

M/(i.,Z ~TuO -270(0 PII. 

-.1'/. oo,,~s'" 1fODI'~ ""'~ "~ .... o 

Pi,. CoOOO/ '" .,,0 rl-.. ~ '.01'" OCSc.ltll'rrot>r "'" .... "' ~ 
I 

Figure 2.10 

Variable Speed Reciprocating Rig 

22 
1) Mechanical frictIon In the crank drive mechanism (Q ); o 

2) PIston ring frictIOn (Qf); 

3) PIston ring leakage (QL); and 

4) HysteresIs In the gas thermal boundary layer (Qh). 

Head developed In these processes IS dISSIpated as follows: 

1) Mechanical frictIon heat IS dISSIpated by the lubricatIon system 

2) PIston ring frictIOn heat IS disspated by the water Jacket 

3) Leakage and hysteresIS heat IS dISSIpated partly by radIatIon and 

convectIon from the cylinder wall and partly by the cooling 

Jacket. 

The above energy flow IS summarized In FIgure 2.11. It IS Important to 

note that the leakage loss (OJ IS really composed of two separate effects: 

1) The phase shift effect (QLp)' caused by a phase shift between the 

volume VariatIOn and the bulk pressure VariatIOn, equally Influences 

every infinItessImal volume of gas In the cyhnder. It may be 

consIdered as a dIstributed volume source WhICh transfers heat to the 

cylinder wall. Part of that heat IS radIated and convested to the ambIent 

and part IS conducted to the water Jacket. 

2) The OrifICe effect (QLo) dISSIpates heat locally at the pIston ring 

by the frictIonal resIstance of the gas leaking through the OrifIce. 

In the test rig, analytIcal consIderatIOns (see AppendIx I) indIcate that the 

phase shift effect IS about 40% smaller than the orifice effect. In an engine, 

however, the SItuatIOn IS completely reversed Since the phase shift effect 

dIsturbs the princIpal process of energy converSIOn and thus accounts for a major 

loss 'IIvhereas the ofIfIce effect IS a relatIvely minor parasItIc loss. 

23 
--Q-h--~~--~----------~ 

---) Qo ------

W Work mput {measured by balance} 

Qo 

Heat dIssIpated by lubricatIon system 

QLp PIston rmg leakage heat by phase-shIft effect 

QLo PIston rmg leakage heat by orIfIce effect 

Qh HysteresIs heat 

Qf PIston rmg frictIOn heat 

Qc Total heat dIssIpated from un cooled cylInder wall 

Qw Total heat dIssIpated by water Jacket 

Figure 2.11 

Energy Flow Diagram for the Piston Ring Test Rig 

24 
For the tests descnbed here the Important fact IS that the heat generated 

by the onfice effect IS dIrectly transmItted to the water jacket smce It IS 

generated locally at the rmg. The heat generated by the phase shift effect 

can be regarded as emanatmg from a dlstnbuted volume source and IS fIrst 

transmItted to the cylmder wall from WhICh a part IS conducted to the water 

jacket and the balance convected and radiated to the surroundmgs. 

The work mput (W) IS dIrectly measured by the balance and the total heat 

dISSIpated by the water jacket (Q ) IS measured by measurement of the water 

w 

flow and the temperature difference ( ~T) between the mlet and outlet of the 

jacket. 

As long as the water jacket temperature and geometry stay constant the 

quantIty (Q Lp + Q h) as well as the fractIon (£) of that quantIty whIch IS 

conducted to the water jacket are uniquely related to the temperature of the 

cylmder wall. 

The relatlOn between (QLp + Q h) and the cylinder wall temperature (Tc) 

was determmed expenmentally by placmg a heatmg element mSlde the statlonary ng cylmder, runnmg water through the jacket, and measurmg the cyhnder 

wall temperature and the water jacket ~T for vanous levels of electnc heat 

mput (Q ). The results of these measurements are shown m FIgure 2.12. In , 

Smce the purpose IS to measure the fnctlOn loss (Q f ) and the leakage 

loss (Q L) of the pIston rmgs, and smce the above measurements YIeld the 

followmg quantItIes: 

Qw == Qf + QLo + £ (QLp + Qh) 

QLp + Qh :: Qvol 

25 

and 
N 

0\ 

Heat (w) 

60 

50 

40 

30 

20 

10 

Amb1ent Temperature: 20°C 

Water Jacket Temperature: l3°C 

o Heat d1ss1pated by water Jacket 

+ Total heat mput 

o +1-----+-----+-----+-----+-----+-----+-----+-----+ 

5 15 25 35 45 55 65 75 85 

Cyhnder Wall Temperature (above amb1ent) (OC) 

Figure 2.12 

Dependence of the volume heat mput and 1tS fractlOn conducted to the water Jacket on the cylmder wall temperature 
we must be able to separate the contributIons of frictIOn, leakage, and 

hystereSIS to the above quantItIes. Such separatIon IS facIlItated by the 

dIfferent mfluence of the speed on each of three losses: 

• PIston rmg frictIOn heat loss per cycle IS mdependent of the speed. 

(see AppendIx II): 

• PIston ring leakage heat loss per cycle IS mversely proportIOnal 

speed (see AppendIx I); 

• HysteresIs heat loss per cycle IS mversely proportIOnal to the 

square root of the speed (see, e.g. Ref. l) 

WIth Q f ' Q Land Q h known, the work mput W, measured by the balance 

YIelds the drIv€, frictIOn (Q ) as follows: 

o 

2.6 - Data ReductIOn for PIston Rmg Performance 

Data WhICh was taken early m the program dId not mclude coolant flow and 

temperature dIfference and dId not mclude cylInder wall temperature. Therefore all the data whIch dId mclude these measurements was used to characterize 

the drive losses (Qo) WhICh was then applIed to the rest of the data. ThIS 

method also mcreased the accuracy smce the work mput data mdicated by the 

scale IS more accurate than the water Jacket data. 

Accordmg to SectIOn 2.5 above the followmg quantItIes were measured: 

Qw ; Qf + QLo + € (QLp + Qh) 

Qyo1 ; QLp + Qh 

E: = 1 - Q/Qyo1 

27 

(2.1) 

(2.2) 

(2.3) 

(2.4) 
SolutIon of the above four equatIons for.Q In terms of the measured o 

quantItIes on the left hand sIde of the equatIons YIelds: 

A drive frictIon coeffICIent (fd> wa~ defined a5 foHows: 

where 

db IS the main bearings radIUS; 

Ap IS the pIston area; and 

P m IS the mean pressure 

(2.5) 

(2.6) 

HydrodynamIc lubricatIOn theory stIpulates a hnear relatIOn between the 

frictIon coeffIcIent f 0 per 2.6 above and the non-dImensIonal parameter S 

defined as follows: 

s = (2.7) 

where 

j.1 IS the dynamIC VIscosIty of the 011; and 

W IS the angular speed 

Since the drive frictIOn IS Independent of the rings used all the available data (34 data points of three dIfferent rings) were used to fit a hnear 

curve (by the least squares method) to the relatIOn betwen f d and S. 

The data and the fItted curve IS shown In FIgure 2.13. Q IS now gIven o 

faIrly accurately as a functIOn of speed (w) and mean pressure (p ). m 

28 
20 

15 

FrIctlon Coefflclent 

10 ~ 

N 

\0 

5 I 

8 

0,.. 

+ 

~T 

A~+ 

-P9 

A 4 0 A 

B A + 

4° 

+ 

-+-

• 

T 

• 

+ 

-r 

-+-

• 

• 

1".5 2 2'.5 3 3:5 il 4·.5 

Speed - Pressure Parameter S = 11 w/Pm 

Figure 2.13 

Drive Coefflclent vs. Speed-Pressure Parameter 

• 1 MPa 

+2 MPa 

/l4 MPa 

05 MPa 
The fItted curve IS: 

SubstItutIon of 2.6 and 2.7 m 2.8 YIelds: 

1l db A P P m 

and rearrangmg: 

til 

+a 

o 

It db A (a p + a 1 J..I til) porn 

ThIs was now applied to all the data mcludmg such measurements where 

coolant and cylmder wall temperature data were not available. The dnve 

(2.8) 

(2.9) 

( 2.10) 

losses (Q ) per 2.10 above was subtracted from the measured work mput (W) to 

o 

YIeld a quantIty (W- Q ) whIch accordmg to SectlOn 2.5 above IS made up of 

o 

the followmg: 

(2.11) 

Least squares were used to fIt the curve on the extreme nght hand sIde of 

2.11 to each set of (W - Q ) quantItIes taken for a partIcular rmg at constant 

o 

mean pressure. 

2.7 - Results 

The results of the performance measurements of the four sets of pIston 

rmgs are presented and dIscussed m thIS sectlOn. Tables 2.1 through 2.4 list 

the vanous losses for all four sets of rmgs at dIfferent speeds and pressures. The data IS further reduced to achIeve a clearer mdIcatIon of trends 

30 
Table 2.1 

Summary of Energy Flow per Cycle - RlIlg 5et tIl 

t-.1ean Pressure (MPa) 

2 3 4 6 

Qf (all speeds) 5.3 I 4.14 1 5.63 

~--------------4---------4---.-------+--------~--------_~I--------_~I---------1 Speed (RPM) 750 I 750 ! 

! 

I Speed (RPM) 

I 

20.7 

6.24 

8.54 

31.10 

1125 

33.35 

7.74 

38.56 

46.30 

1125 

I i QL i 13.80 22.23 I 

Qd i 8.54 10.12 ! 

750 

45.3 

I 9.21 

! 

50.97 

60.19 

I 1125 

30.20 

i 11.70 

~ ! 17.92 27.80 I 

\V I ! 26.l~6 37.02 I 47.99 1 

-----~i---------f--------+--------+--------,------~l 

35.70 

Speed (RPM) I 1500 1500 I 1500! I I I 

QL I 1 10.35 I 16.68 I 22.65 : 

Qd I 10.63 I 12.38 I i 14.62 I 

Qw I 14.51 I 21.79 I : 28.37 

W I 75.13 I 34.17 I I 42.99 

i I I I 

~---s-pe-e-d-{-R-P-t-.-1)--~II·--------~;--------~I~----------r--------~i----------;-------~ 

I I I I I I 

QL ' I I I I I , 

·l ____ ~_: ______ . __ ~ ______ ~ __ . ____ ~~ ______ ~ ______ ~ ________ ~ _____ ~ Where: Of = Piston Rmg FrictIOn - Joule'> 

QL .: Piston Rmg Pha::,e <)hJ ft plus OrifICe Leakage Effectc; - Joulec; 

Qd -: Crank Drive FrictIOn - Joules 

Q = Heat Dlsslpdted by \Vater Jacket - Joules 

w 

\V - Merhamcal Encrp,y Input - JOllies 

31 
Table 2.2 

Summary of Energy Flow per' Cycle - Rmg Set 114 

Qf (all speeds) .47 

Speed (RPM) 794 

QL 6.35 

Qd 5.53 

Qw 

6.70 

W 12.24 

Speed (RPM) 1197 

QL 4.21 

Qd 7.89 

Qw 

5.00 

W 12.90 

Speed (RPM) 1810 

QL 2.79 

Qd 11.46 

Qw 

1.22 

12.68 W 

Speed (RPM) 2210 

QL 2.28 

Qd 14.14 

Qw 

4.38 

W 18.52 

2 

1.77 

790 

8.87 

6.63 

10.13 

16.76 

1205 

5.81 

8.77 

7.76 

16.54 

1805 

3.88 

12.49 

5.27 

17.75 

2200 

3.19 

14.62 

5.48 

20.10 

Mean Pressure (MPa) 

3 4 

4.47 

785 

11.10 

7.99 

15.37 

23.37 

1196 

7.28 

10.33 

12.26 

22.60 

1805 

4.83 

14.03 

8.90 

22.92 

2209 

3.94 

16.18 

9.60 

25.79 

Where: Qf = PIston Ring FrIctIOn - Joules 

5 

6.93 

805 

10.10 

9.14 

16.77 

25.90 

1194 

6.79 

11.70 

13.70 

25.35 

1815 

4.46 

14.62 

11.17 

25.79 

2212 

3.66 

17.30 

10.69 

I 27.99 

QL = PIston Rmg Phase ShIft plus OrIfICe Leakage Effects - Joules 

Qd = Crank DrIve FrIctIon - Joules 

Q = Heat DISSIpated by Water Jacket - Joules 

w 

W = Mechanical Eneq~y Input - Joules 

32 

6 

I 
Qf (all speeds) 

Speed (RPM) 

QL 

Qd 

Qw 

W 

Speed (RPM) 

QL 

Qd 

Qw 

W 

Speed (RPM) 

QL 

Qd 

Qw 

W 

Table 2.3 

Summary of Energy Flow per Cycle - Rmg Set 112 

2.86 

797 

5.43 

5.61 

8.28 

13.89 

1196 

3.61 

7.95 

6.47 

14.42 

I 

2 

Mean Pressure U..1Pa) 

3 4 

3.95 6.94 7.89 

790 812 810 

8.69 8.68 10.26 

6.34 7.31 8.58 

12.51 15.18 17.88 

18.85 22.49 26.46 

1100 1174 I 1184 

6.24 6.00 7.02 

8.09 9.36 10.33 

10.43 15.18 15.46 

18.52 23.15 25.79 

1830 2012 2025 

3.75 3.50 4.11 

12.58 I 14.62 15.60 

7.71 I 10.51 I 11.90 

20.28 25.13 27.56 

5 

12.46 

810 

7.37 

9.50 

19.60 

I 29.10 

I 1213 

I 

4.92 

11.70 

17.84 I 

I 29.54 I 

! 2031 I 

I 

I 2.94 I 

, 15.84 I 

1 15.69 

I 

I 31.53 

I 

I 

I 

, 

I 

I 

I 

I 

I , 

I 

I 

, 

I 

I 

I 

I 

i 

I 

I 

I 

, 

I 

I 

I 

I 

I I I 

I 221::1· I 

Speed (RPM) 2190 2220 2197 I I I I 

QL I 3.13 3.17 3.78 I 2.69 

Qd 14.62 15.94 16.57 I 17.55 

Qw 

6.98 9.64 11.43 

I 

I 14.64 I 

W 21.61 25.57 27.99 I 32.20 ___ L ____ 

Where: Qf = Pl~ton RIn~ FrictIon - JOllies 

QL :: PIston Ring Phase ShIft plus Or,! Ice Ledh.agc> Ef kn,> - _~oule<; 

Qd -= Crank Drive FrIctlOn - Joules 

Q = Heat DISsIpated by \\ atE'r Jacket - Joules 

v,' 

\V -= \1cc.hanlcal Eneq:;v Input - Joules 

33 

J. 

I I 

I --, I 

I 

I 

I 

I 

---l 

1 

I 

I 

I 

I 

I 

i 

I 

.~ I 

I 

I 

I 

. 
Table 2.4 

Summary of Energy Flow per Cycle - Rmg Set 113 

2 

Qf (all speeds) 2.44 5.10 

Speed (RPl\1) 800 800 

QL 4.04 5.27 

Qd 5.61 6.24 

Qw 

6.52 10.30 

\V 12.13 16.54 

Speed (RPM) 1200 1200 

QL 2.69 3.51 

Qd 7.97 8.87 

Qw 5.04 8.77 

W 13.00 17.64 

Speed (RPM) 2000 2000 

QL 1.62 2.11 

Qd 12.67 13.45 

Qw 

4.30 7.27 

17.00 20.72 \V I 

Speed (RPM) 2225 2225 

QL 1.45 1.89 

Qd 14.14 I 14.96 

Qw 

3.70 6.86 

W 17.90 21.83 

r..1ean Pressure (MPa) 

3 4 

6.81 9.11 

800 800 

6.15 6.22 

7.60 7.80 

12.68 15.13 

20.28 22.98 

1200 1200 

4.10 lL.15 

9.43 10.43 

11.50 13.60 

20.94 24.00 

2000 2000 

2.46 I 2.49 

14.55 14.62 

9.26 12.27 

23.81 26.90 

2225 2225 

2.21 2.24 

15.99 16.57 

8.72 10.54 

24.69 27.12 

Where: Qf -= Piston Rmp' Friction - Joules 

5 

14.5 

800 

2.47 

9.26 

16.97 

26.23 

1200 

1.65 

10.97 

16.15 

27.12 

I 

QL = Pl~ton Ring Phase Shift plus Orlflc.e Leakage Effects - Joule,> 

Qd = Crank Drive FrictIOn - Joules 

Q = Heat DISSipated by Water Jacket - Joules 

w 

W - \1P,hantct! F.ncr~y Inpllt - Joules 

34 

6 

12.72 

800 

7.677 

9.80 

19.97 

29.76 

I 

11200 

I 5.12 

I 11.70 

18.73 

30.29 

. 

2225 

2.76 

18.29 

15.01 

33.29 
as follows: 

2.7.1 FrictIon - The frictIon data was reduced USing equatIon II-8 from 

AppendIX II to glean the frictIOn coeffIcIent f: 

f 

2 1t D S h Pm n (t.v/V m) 

If the frictIOn mechanism IS Indeed the one modelled In AppendIx II then 

the frictIOn coeffICIent f IS only a functIon of the matenal. 

FIgure 2.14 shows the fnctIOn coeffeClent of all rings plotted versus the 

mean pressure p . There IS a substantIal spread of the data about an average m 

value of 0.17 whIch IS somewhat hIgher than that (0.15) commonly specIfIed In 

the lIterature. The spread of the data appears to be random and does not show 

any partIcular trend. Both the data spread and the slIghtly hIgher average 

value may be explained by the fact that the hysteresIs losses, being of the 

same order of magnitude as the error In the measurement, are partIally Included 

In the frictIon and leakage losses In a random manner. Since leakage loss In 

the rig IS much hIgher than the frictIOn loss the hysteresIs error Influence 

on the latter IS much hIgher than on the former. The conclUSIOn IS that the 

frictIOn loss IS proportIOnal to the aXIal thIckness (h) of the rings and, more 

generally, that the frictIon mechanism IS correctly modelled In AppendIX II. 

2.7.2 Leakage - The leakage data was reduced to elIminate the mfluence 

of speed USing equatIon 

35 
0.5 

0.4 

FnctlOn CoefflClent 

0.3+ • • Ring Set 111 

+ i- Ring Set 112 

+ D. Ring Set 113 

VI 

0.21 

0\ 

;-

• 0 Ring Set IIl~ 

+ ;- i 

A 

A fA & • A 

0 

• 0.1 

I 

0 

o~I ____ ~ __ ~----~----r---~----~ 

o 2 3 4 5 

Pressure (MPa) 

6 

Figure 2.14 

Plston Rmg Coefficlent vs. Mean Pressure 
The "leakage coefficIent" CL is a monotonically increasing functIon of the 

mean pressure (linear for laminar restrictIOn) the slope of WhICh IS Inversely 

proportIOnal to the restrictIOn of the ring. PhysIcally, C L IS proportIOnal to 

the leakage power loss and ItS Units are those of power: 

FIgure 2.15 shows the behavIOr of the coeffICIent CL wIth pressure for the 

four sets of rings. 

The most striking trend exhIbIted by FIgure 2.15 IS the large dIfference 

between the leakage of Ring Set III (the set wIth untreated surfaces of the step 

JOints) and that of all other ring sets. 

ThIS clearly demonstrated that It IS the quallty of the step JOint WhICh 

exerts the most Influence on the leakage characteristIcs of pIston rings. 

Comparison of the data of Ring Sets If 2, 3 and 4 among themselves indIcate the 

follOWing: 

• Ring set II 3 WhICh has the largest aXial thIckness exhIbIts somewhat 

lower leakage than that of the other two sets (about 30% lower). 

• Ring sets If 2 and 4 exhIbIt practIcally the same leakage 

• Except for the reading at 5 MPa the data shows relatIvely llttle 

scatter and generally exhIbIts a llnear trend In the pressure Influence modifIed by a certain "flattening" at hIgher pressure indIcating 

the presence of some turbulent effects on the leakage mechanism. 

2.7.3 Comparison of OrifICe Effects to Phase Shift Effects - Strictly 

theoretIcal conSIderatIOn (see AppendIX I) indIcate that, for both laminar and 

37 
s 

L1 

T 

• 

3 t • Ring Set III 

• + Ring Set 112 

C,.N 

Leakage CoeffICIent (kW) I ~ Ring Set 113 

00 

2 ...L 0 Ring Set 114 

• 

~ 0 

0 + + fJ. 

~ .1 II A 

.d A 

I 0 I I I I I 

0 1 2 3 4 5 6 

Pressure (MPa) 

Figure 2.15 

PIston Ring Leakage CoeffICIent vs. Mean Pressure 
Table 2.5 

RatIOs of OrifICe to Phase-ShIft Leakage Effects 

Mean Pressure Nommal Speed 

(MPa) ( RPM) 112 114 

1 800 1.09 1.37 

1 1200 1.51 0.83 

1 1800 0.73 0.81 

1 2200 1.05 1.69 

2 800 2.05 1.54 

2 1200 1.78 1.37 

2 1800 1.68 1.11 

2 2200 2.41 1.47 

4 800 1.33 2.07 

4 1200 1.20 1.90 

4 1800 1.17 1.84 

4 2200 1.52 1.69 

5 800 1.74 1.06 

5 1200 1.72 1.06 

5 1800 1.39 1.15 

5 2200 2.05 1.23 

Average QLo/QLp 1.53 1.38 

Combmed A veraBe 1.46 

39 
turbulent resIstance, the ratIO of the OrifICe effect power loss to the phaseshift effect power loss (QLo/QLp) IS equal to the polytropIc constant (n) of 

the compressIon and expansIon processes. 

For such rmgs where coolant and cyhnder wall temperature were available 

(sets II 2 and 4) the data was reduced per SectIon 2.5 above to separate the 

above two leakage effects and mdicate the ratIO. 

The results, tabulated m Table 2.5, show that, withm the error of the 

measurement, the average ratIo 1.46 0.38 and 1.53 for Rmg Sets II 4 and 2 

respectIvely). 

The polytropIc constant is expected to be somewhat lower than the 1.67 

characterizing purely adIabatIc processes for the monoatomiC hehum, smce some 

heat exchange wIth the ambIence does occur. The average result of 1.46 IS 

entIrely reasonable and mdicates that the data and the data reductIOn method 

do reflect the real physcial processes WhICh occur m the testmg. 

2.7.4 Accuracy and Data Spread - Some mdicatIOn of the accuracy of the 

data IS afforded by comparmg the actual data pomts mdicatmg the total 

frictIOn leakage and hysteresIs loss (Q t) to the values provIded by the 

ac 

curve fitted to the data (Q fIt = C f + C l / w). The ratIo (Q /Qf ) IS J- act It 

plotted versus speed and pressure for each set of rmgs and shown m FIgures 

2.16 through 2.19. 

The followmg observatIons may be made: 

• By and large, no trend can be dIstmgUished m the mfluence of 

pressure and speed on the data spread. 

• The data spread of Rmg Set II 1 IS extemely small (± 1% max.). ThIS 

IS because the total loss (partIcularly leakage) IS so large. 

Results for Rmg Set III are thus very accurate and reliable. 

40 
.j::o. 

I--' 

1. 01 

1.008 

1 .006 

1.004 

1.0021 

0.998 t 

0.996 

0.994 

0.992 

0.99 

700 

• 

A 

1-

(Measured Heat)/(Fltted Heat) Values 

A 

• 2 MPa 

• 

+ 4 MPa 

A 6 MPa 

ff 

A 

+ 

+ 

ROO 900 1000 1100 l~no 1300 1400 1500 

Speed (RPM) 

Figure 2.16 

Quality of Curve FIt to Measured Data - Rmg Set If I 
1.1 

(Measured Heat)/(Fltted Heat) Values 

A 

1.05+ 

• 1 MPa 

0 

)( + 2 MPa + X 

A 3 MPa 

.A 

.". 

10 0 4 MPa 

N 

~ + 5 MPa 0 )( 

6 )( 

0.951 A 

0.9 

00 800 1000 1200 1400 1600 1800 2000 2200 2400 

Speed (RPM) 

Figure 2.17 

QualIty of Curve FIt to Measured Data - Rmg Set 112 
1.1 

0.9 

+ 

(Measured Heat)/(Fltted Heat) Values 

o 

• 

600 800 1000 1260 1400 1600 1800 2000 2200 2400 

Speed (RPM) 

Figure 2.18 

QualIty of Curve FIt to Measured Data - Rmg Set 113 
SECTION 3 - RECIPROCATING ROD SEAL 

3.1 IntroductIOn and Summary 

The StIrlmg engme concept which has been developed at StIrlIng Thermal 

Motors, Inc. mvolves a van able swash plate dnve. Power control IS accomplIshed by means of pIston stroke VariatIOn affected by changmg the swashplate 

angle. The mean cycle pressure m the engme thus remams constant at all 

operatmg condItIOns. 

ThIS makes It possIble to mamtain the crankcase at mean cycle pressure 

and thus relax the functIOnal requIrements of the reciprocatmg rod seal so 

that they are reduced to the followmg two: 

• Isolatmg the varymg cycle pressure from the constant crankcase 

pressure; and 

• preventmg 011, present m the crankcase, from penetratmg the 

thermodynamIc cycles. 

The fIrst of these two functions IS readIly accomplIshed WIth a cap seal 

WhICh IS a Rulon (or SImIlar thermoplastIc material) rmg backed up by an O-rmg 

mounted around the pIston rod. 

Regardmg the preventIOn of crankcase 011 from penetratmg mto the 

cycles, prevIOus expenence WIth asymmetrically loaded scrapers made of babbIt, 

or a SImIlar soft metal, YIelded very good results provIded there was no 

clearance between the scraper and the rod. 

Lateral motIOn of the rod, due to normal manufacturmg tolerances 

effectmg the cross-head clearance, necessarily causes clearance between the 

scraper and the rod to appear smce the rod wIll wear down the babbit. ThIS 

WIll cause the scraper to cease functIOnmg properly after a short tIme of 

46 
• Rmg Sets 1/ 2 and 3 exhIbIt sImIlar data spread of about 6% 

• Rmg Set /I 4 exhIbIt the largest spread of 14% and also show an 

unexpected mfluence of pressure (a steady mcrease m the spread 

from 7% at IMPa to 14% at 3 MPa and abrupt recductIOn to 2% at 

4MPa). Data for thIS rmg set IS the least accurate and relIable and 

should be consIdered merely as prelIm mary mdicatIOn of performance. 

45 
1.8 

1.6 

1.4 

1.2+ 

-+>- I I 

-+>- t 

0.8 

0.6 

0.4 

600 800 

(Measured Heat)/(Fltted Heat) Values 

• 

• f 

~ d l' Ii 

+ • 

t 

• 

1000 1200 1400 1600 1800 2COO 22lJO 

Figure 2.19 

Quallty of Curve FIt to Measured Data - Rmg Set 114 

-r 

• 

+ 

A 

0 

2400 

Speed (RPM) 

1 MPa 

2 MPa 

4 MPa 

5 MPa 
operatIon. 

To address this problem, StirlIng Thermal Motors, Inc. has designed a 

complIant housing. The scraper IS mounted to the complIant housing and IS 

forced against the rod by a crown spring also mounted to the complIant housing. 

The entIre assembly can move laterally wIth the pIston rod wIthout any 

apprecIable force exerted on the scraper due to such motIOn. 

The cap seal IS also mounted In the complIant housing and follows the 

lateral motIon of the rod. 

The comphant housing IS made of nylon. Its convoluted geometry makes It 

rigId aXially and complIant laterally. The design IS described In detaIl In 

SectIOn 3.2 below. 

The experimental assessment of the rod seal, hereunder, was aimed at the 

following two obJectives: 

• Assess seal endurance under design condItIOns, VIZ. wIth no pressure 

dIfference across the seal but with an appreCIable lateral motIOn of 

the rod; and 

• Assess much of the applIcabilIty of the seal as a containment seal 

for the gas as well as an 011 scraper. 

Endurance testing was performed at constant speed of 1800 RPM with 

atmospheric pressure on both sides of the seal and with lateral motIOn of the 

rod of an amplItude of 60 llm. By July 8, 1985, the seal has completed 6750 

hours of operatIOn without any faIlure and with no degradatIOn of performance. 

During the entIre test, frequent Visual inSpectIOns discovered no 011 at all 

above the seal. (There was one problem at about 400 hours which was specIfIC 

to the test rig and does not indIcate an Inherent seal problem. This IS 

deSCribed In SectIOn 3.3 below). 

47 
These results are very encouragmg and mdicate the adequacy of the rod 

seal and the compliant housmg to engmes wIth a pressurized crankcase. 

As a contamment seal, with mean pressure dIfference across, the rod seal 

wIth the compliant housing dId not perform very well. The followmg problems 

were IdentifIed: 

• The low tenstle strength of the nylon housmg caused It to break 

under a pressure dIfference of 8 MPa across It. 

• Skewed deformation of the nylon housmg under pressure, due to some 

devIatlOn from aXial symmetry, mduced plastIc deformatIon of the 

soft babbit scraper. When the pressure was reduced thIS plastIc 

deformatIon gave rise to some geometric mcompatibility WhICh 

prevented the seal from sealing 011. 

In order to address these problems It IS felt that a new compliant housmg 

design IS needed, such that WIll retam the compliance but WIll add SIgnifIcant 

tenstle strength. In additlOn, the elastIC limIt of the scraper element Itself 

should be mcreased. ThIS may be accomplished by usmg a stronger metal, 

pOSSIbly plated wIth a layer of softer metal. 

3.2 Rod Seal DeSIgn 

An assembly drawmg of the scraper rod seal IS shown m FIgure 3.1. The 

scraper (1) IS made of tm babblt (ASTM 3) and IS loaded agamst the rod by a 

crown sprmg (2) whIch exerts an even pressure on the sealing edge through a 

square rubber rmg (3). The square rmg (3) IS retamed between the scraper 

housmg (4) and the spacer (5), and thus also seals the OD of the scraper (1). 

All the above components are mounted in the scraper body (6) whIch threads onto 

the com pilant nylon housmg (7) WhICh also con tams the cap seal (8). The 

48 
-6 

I 

Figure 3.1 

-7 

Scraper Rod Seal Assembly 

1) Scraper (ASTM3 tin babblt) 

2) Crown sprIng 

3) Square rubber ring 

4) Scraper housing 

5) Spacer 

6) Scraper body 

7) ComplIant housing (Nylon) 

8} Cap seal (Rulon LD) 

9} Assembly pIlot (replacing pIston rod) 

49 
complIant housing (7), In turn 1S threaded to the engine block. Photographs of 

the assembly and two stages of d1sassembly are shown In F1gures 3.2, 3.3 and 

3.4. The scraper 1tself 1S shown In F1gure 3.5, and the crown spring In 

F1gure 3.6. 

3.3 DeSCriptIOn of the Tests 

The endurance test was performed on a rec1procatIng rig, shown In F1gure 

3.7, wh1ch was operated w1thout a p1ston In order to deprive the rod from 

, 

gU1dance. The space above the scraper assembly was open to the atmosphere and 

equ1pped, nght above the seal, with inSpectIOn ports. The space below the 

seal, on the crankcase s1de, 1S also at atmospheric pressure and 1S equ1pped, 

right under the seal, w1th a plex1g1ass-covered inSpectIOn port through wh1ch 

011 splashing against the bottom of the scraper housing could be seen. 

Clearances In the cross head and the bearings were such that the lateral 

mot10n of the rod was measured (stat1cally, with an Ind1cator) to be of an 

amplItude of 60 ~ '1"1. 

The ng was driven at 1800 RPM w1th a constant speed electriC motor and 

operated w1thout problems from the outset for about 400 hours. At that pOint a 

redd1sh OIly substance appeared on top of the complIant housing. InspectIOn 

revealed that the substance was a m1xture of 011 and Rulon dust from the cap 

seal. In fact, 1t was found that the scraper area was choked w1th Rulon dust. 

Upon further InVest1gatIOn 1t was found that the cap seal was wearing 

excess1vely Since the roughness of the rod was seriously out of specIf1catIOn 

(4-5 m1cro Inches compared to the specIf1ed 14-18 m1cro Inches). Also the 

part1cular OrientatIOn of the rod seal w1th the scraper below the cap seal 

caused the wear part1cles to accumulate at the scraper area. 

The scraper assembly was cleaned and reinstalled (wIth the same scraper) 

so 
Sl 

Figure 3.2 

Scraper rod seal assembly 

3 .. 3 

Scraper rod seal, partially 

disassembled. (From right: 

capseal, compliant housing, 

scraper body assembly, crown 

spring, scraper housing) 

Scraper rod seal components 

(Top, from right: Scraper 

body, scraper, spacer, square 

rubber ring) 
Figure 3.5 

Scraper 

Figure 3.6 

Crown Spring 

52 
- - 7 

- 6 

4 

5--t--~+-

2 

9 

Figure 3.7 

Rod Seal Endurance Test Rig, Schematic 

(1) Crank shaft (2) Cross heads (3) PlexIglas covered mspectIOn port 

(5) Scraper rod seal (6) PIston rod (7) Cylmder (8) CouplIng 

(9) Constant speed AC motor 

53 
and was operatmg perfectly agam, from the outset. 

By July 8, the scraper had completed a total of 6750 running hours. 

The endurance test strongly mdIcates that the scraper seal m Its 

complIant housing WIll operate very well In an engine WIth a pressurIzed 

crankcase. It IS recommended that thIS seal be tested further under more 

realIstIC condItIons and that, perhaps, some quantItatIve data relatIve to 

frIctIon, wear and 011 permeatIon be taken. 

Testing of the scraper assembly as a containment seal was performed on the 

variable speed rIg deSCrIbed m SectIon 2.4 above. Indeed, It was the scraper 

assembly, actmg as a contamment seal, which allowed presSUrIZatIOn of the rIg 

and performing the pIston rIng tests. 

The fIrst problem was the lImIted tensIle strength of the complIant 

housmg: It broke at mean pressure value of 8 MPa. ThIS happened early In the 

program and for the balance thereof no attempt was made to go to mean pressure 

values hIgher than 5 MPa. 

At pressures below 5 MPa the follOWing trend was observed: starting at low 

pressure and Increasing the pressure monotonIcally In steps, the seal performed 

rather well, contammg the gas pressure and preventmg 011 mIgratIOn. Upon 

reversmg the process and going down WIth the pressure the seal continued to 

contain the gas pressure, but developed 011 leakage. 

Further inVestIgatIOn, by pressurIzing the complIant hOUSing m a fIxture, 

revealed that under pressure the hOUSing had undergone a skewed, off-aXIS 

deformatIon. It IS hypotheSIzed that such deformatIon, caused by slIght 

deVIatIon from aXIal symmetry, Induced plastIC deformation In the soft babbIt 

scraper. Upon redUCing the pressure, the geometry of the scraper no longer 

matched that of the rod and the scraper could no longer prevent 011 mIgratIon. 

Consequently, for the scraper assembly to functIon adequately WIth a 

54 
pressure difference across it the following should be done: 

• The compliant housing should be redesigned USing different materials 

and different geometry so that itS compliance is retained and itS 

tenSile strength greatly Increased. 

• The scraper itself should be made of a different matenal with an 

elastiC limit much greater than that of ASTM 3. Any such material is 

likely to be too hard and will not conform well at the contact area 

With the rod. Plating the scraper With a layer of soft metal Will 

solve thiS problem. 

55 
SECTION 4 - HEAT PIPES 

4.1 IntroductiOn and Summary 

The extensIve use of heat pIpes, mdeed, theIr fundamental role m the 

StIrhng engme concept of StIrhng Thermal Motors, Inc., called attentiOn to 

the fact that current hIgh-temperature sodIum heat pIpe practIce renders thIS 

sImple devIce somewhat ImpractIcal. 

The current practIce calls for complex and expensIve fabricatiOn and 

handlmg procedures reflectmg concerns that are, perhaps, JustifIable m the 

nuclear power plant apphcatlOn (the major source of thIS technology) but seem 

undue m the StIrhng engme apphcatlOn. 

The purpose of the effort reported here was, generally, to demonstrate 

that sodIum heat pIpes for engll1e apphcatiOn are sImple, practIcal devIces. 

In partIcular, the testmg program was set to evaluate the followmg: 

• Use of sImple fIllmg procedures mvolvmg commercIal grade sodIum, 

undlsttlled and handled In atmosphere 

• Use of getters to avoId furnace degassmg and mcrease the tolerance 

for aIr leakage 

• Use of mechamcal couphng to Jom heat pIpe sectlOns. 

By July 8, 1985, a total of SIX heat pIpes have accumulated about 46,000 

operatmg hours. SIX other heat pIpes dId not operate well at all. The 

unquahfied success of the former and the failure of the latter clearly pomts 

at the followmg conclusions: 

• SImple and crude fIllmg procedures mvolvmg msertlOn of sohd 

pIeces of sodIum, handled m aIr, mto the open heat pIpe prior to 

weldmg a cap and evacuatmg can be adequate. In addItiOn, only 

56 
70 cm 

36 cm 

Valve 

11---1/4" extensIon tube 

Condenser cap 

~ ........ ~~-t-r- Getter container hook 

J---HI--- Getter container 

Figure 4.1 

Heat pIpe body 

H--- 1 1/2" x 0.049" tube 

CoIl spring 

....... -- WIck 

Evaporator cap 

Typical Heat Pipe, Schematic 

59 
ready for normal operatIOn. 

FIgure 4-.2 shows, schematIcally, the vanous heat pIpe confIguratIOns 

used. Table 4-.1 summanzes some of theIr charactenstics. 

Heat pIpes 112 and 113 are the base-hne confIguratIOn and have no getters. 

Heat pIpes 114- and 115 have getters con tamed m a perforated tube suspended m 

the condenser area from a hook attached to the condenser cap. ThIS confIguratIOn proved unsuccessful smce the calclUm (WhICh IS one of the getters) melted 

at 830°C and leaked out through the holes m the getter tube, contammatmg 

the wIck. The temperature was, at the tIme, controlled manually at about 800°C, 

but, due to the diffIculty of manual control, fluctuatIOns occaSIOnally brought 

the temperature above the meltmg pomt of the calclUm. When the temperature 

would swmg back the calclUm would sohdlfy m the wIck and cause overheatmg 

of the contammated regIOn. 

To address thIS problem heat pIpes 116 and 118 were c.onstructed wIth the 

getter contamed m a "bucket" wIth a loosly fItted cover so that even If the 

calclUm melted 1t would remam m the contamer and not contammate the wick. 

Hea t pipes Ill, 119, and 1110 were bullt with a getter tube extendmg through 

the entIre length of the heat pipe and with no valve. ThiS was done m order to 

assess the possIbIhty of aVOIdmg the evacuatIon operatIOn prIOr to start-up. 

The mtentIOn was to start the heat pIpe up when full of atmosphenc aIr and 

rely entIrely on the getters to evacuate It by absorbmg all the non-condensable 

gases. ThIS concept dId not prove successful and, due to lack of tIme, was not 

mvestigated further. 

Heat pIpes 1111 and 1112 were used to assess the feasibIhty of usmg the 

Conoseal couplmgs (avaIlable commerCIally from AeroqUlp CorporatIOn) shown m 

Figure 4-.3. Smce one of the reqUIrements of the couplmg IS to prOVIde caplllary connectIon, these heat pipes were deSIgned to operate agamst graVIty, 

60 
I 

I 

~I I 

~I I 

2, 3 4, 5 6, 8 

I~ 

7, 9, 10 11, 12 13 

Figure 4.2 

Configurations of the various heat pipes tested. Heat pipe numbers 

are indicated under the configuration schematics. 

61 
Table' 4.1 

Some Charactenstlcs of Heat Pipes Tested 

Heat Pipe Na Ca La Zr Spring Mesh 

II (31655) 

2 16 0 0 0 NI-Cr 2 x 165 m, .002" w 

3 16 0 0 0 NI-Cr 2 x 165 m, .002" w 

4 16 6 6 0 NI-Cr 2 x 165 m, .002" w 

5 16 6 6 0 NI-Cr 2 x 165 m, .002" w i 

I 

6 40 6.5 17.5 0 NI-Cr 2 x 165 m, .002" w 

7 40 13.5 19 0 NI-Cr 2 x 165 m, .002" w 

8 40 8 19 0 30255 2 x 165 m, .002" w 

9 40 14.5 14.5 5.2 30255 2 x 165 m, .002" w 

10 40 68 0 0 30255 2 x 165 m, .002" w 

11 40 0 0 9 30255 3 x 165 m, .002" w 

12 40 0 0 10 30255 2 x 55 m, .004" w 

+ 2 x 300 m, .OOI2"w 

13 40 0 0 13 30255 2 x 165 m, .002" w 

Note: The weIghts of Na, Ca, La and Zr are In grams 

62 
Figure 4.3 

Medium Weight Conoseal Joint with T -bolt Quick Coupler 

63 
wIth the evaporator above the condenser, so that caplllary pressure IS relled 

upon to return the condensate to the evaporator. Consequently, the wIck of heat 

pIpes 1111 and 1112 extends mto the condenser sectIon as well. The capIllary 

connectIon through the couplmgs IS achieved as shown m Figure 4.4. 

Heat pIpe 1113 IS IdentIcal to 1112 except that It has no couplmg. It was 

mtended to provide a companson to the performance of heat pIpe 1112, but It 

faded qUickly due to electnc arcmg m the tube furnace WhICh was used to 

heat ItS evaporator. 

Heat pIpes 1112 and 1113 also con tamed some loose ZIrCOniUm m the condensate path. The ZIrCOniUm was mtended to catallze the dIssocIatIOn of corrOSIve 

Na 0 and bmd the free oxygen. No analysIs of heat pIpe samples has been 

performed to confIrm the effectIveness of the ZIrCOniUm. 

In summary, heat pIpes 112, 113, 116, 118, 1111, and 1112 were useful confIguratIOns whIch were subjected to long term testmg as described m the followmg 

SectIon 4.3. 

4.3 DescnptIOn of the Tests and TheIr Results 

The heat pIpe test stand IS shown m FIgure 4.5. Heat was provided to the 

evaporators by electnc tube furnaces With maXImum capacIty of 1.5 kW. Two 

thermocouples were attached to the 00 of each heat pIpe, one m the evaporator 

and one m the condenser area. These were used to mOnitor and control the 

evaporator temperature, at fIrst manually and later automatIcally, usmg temperature controllers, at 800°C. 

Heat was dISSIpated from the condensers by radIatIOn to the surroundmgs. 

The total heat transport was determmed by the amount of msulatIOn around the 

condensers. Most of the tests were performed at heat transport rate of about 

0.75-1.0 kW. 

64 
Figure 4.4 

CapiJIary Connection through Conoseal Couplings 

1) Heat pIpe body 

2) Wick 

3) Male flange 

4) Female flange 

5) Seahng element 

65 
Figure 4.5 

Heat Pipe Test Stand 

(Shown, from right, Heat Pipes #12, #8, #11, #3) 

66 
Heat pIpes 113 and 118, the former wIthout getter and the latter wIth, have 

operated contmuously from start-up and are stlll operatmg. On July 8, 1985, 

each of these heat pIpes had completed 12,000 hours of problem-free operatIon. 

no difference between the two heat pIpes was eVIdent. 

Heat pIpes 112 and 116 operated wIth no problems until each accumulated 2000 

hours. At that pomt, m order to assess the tolerance of the heat pIpes to aIr 

leakage, metered doses of aIr were mtroduced mto each heat pIpe. Each dose 

was of 2.0 Ncc of aIr. 

Heat pIpe 116 (wIth getters) took about 15 mmutes to recover from each 

dose of aIr, that IS to absorb and bmd 2.0 Ncc of aIr. SurprIsmgly, heat pIpe 

112 (wIthout getter) also recovered and absorbed the aIr, but at a much slower 

rate, of about 15-20 hours per dose of 2.0 Ncc. 

It appears that m the absense of getters the sodIum Itself acts as a 

getter forming OXIdes, nitrides and carbonates, and can handle leaks of up to 

0.1 Ncc/hour untIl all the excess sodIum Inventory is consumed. 

During the next 1000 hours of operatIon, dOSing contmued untll a total of 

58 doses (about 120 Ncc of aIr) were Introduced Into each heat pIpe. At that 

pomt, WIth both heat pIpes stIll operatmg perfectly, the test was stopped and 

heat pIpe 112 was sent out for metalographic inSpectIOn. 

By July 8, 1985, heat pIpe 112 had been inspected and found to be In 

rather good conditIon, WIth only slIght eVIdence of corrosion and element 

transport. The metalographic inSpectIon report IS reproduced m AppendIX III. 

Heat pIpes 1111 and 1112 are bemg tested m order to evaluate the Conoseal 

couplmg. Heat pIpe 1111 IS continuously cycled between 800°C and about 150°C, 

WIth each cycle taking 2 hours. FIgure 4.6 shows a close-up of the couplIng 

67 
Figure 4.6 

Conoseal Coupling in Operation 

(Heat Pipe 1111 at 800°C) 

68 
operating at 800°C. By July 8, 1985, heat pIpe 1111 had completed 4200 cycles 

and was still operating. 

Heat pipe 1112 was operated continuously against gravIty to assess the 

capIllary connectIon through the coupling. The maXImum capIllary pressure was 

determined experimentally to correspond to a maXimum pumping height of 32 cm. 

The heat pIpe was Installed at an angle with the evaporator 32 cm above the 

condensor (right In FIgure 4.5). Any degradatIOn of the performance would Immediately be detected by a general drop In temperature Since the controlling 

thermocouple IS at the hIghest pOint of the evaporator. By July 8, 1985, heat 

pipe 1112 had completed 8000 hours of operation With no problems and no degradatiOn of performance. 

69 
SECTION 5 - SPLIT SHAFT COMPONENT DESIGN 

5.1 - IntroductiOn and Summary 

A swash plate drive for a St1rlmg engme 1S characterized by a heavy 

bendmg moment actmg on the mam shaft. As a result, the unsupported ends of 

the shaft tend to deflect apprec1ably. 

In a swashplate engme w1th pressurized crankcase and a rotatmg shaft 

seal contammg the crankcase pressure, the deflectiOn of the shaft at the 

locatiOn of the rotatmg seal may be more than the amount that the seal can 

tolerate w1thout affectmg hfe and leakage rate. Therefore, a means 1S needed 

to ehmmate shaft runout at the rotatmg seal. 

The problem may be specif1cally 1llustrated by the orIgmal des1gn of the 

STM4-120 variable swashplate St1rhng engme of St1rhng Thermal Motors, Inc., 

shown m F1gure 5.1. 

The mam shaft (4) 1S supported by two closely spaced spherical roller 

bearings (6) and (9), to mm1m1ze deflectiOn due to the bendmg moment apphed 

by the swashplate (8), and a tapered roller bearing (2) to support the pressure 

mduced thrust load and mm1m1ze shaft runout at the rotatmg seal (1). A 

substant1al deflectiOn slope at the rotatmg shaft seal 0) may adversely 

effect the hfe and performance of the seal. 

A solutiOn should be found w1thout mcreasmg the load (thus reducmg the 

hfe) of the mam bearings. In the specif1c case of the STM4-120 engme the 

problem 1S comphcated by the fact that a narrow-clearance bushmg (not shown) 

at the front end of the shaft 1S requ1red for lubrIcatiOn and hydrauhc flu1d 

management. These requ1re that the runout at that end of the shaft 1S also low. 

The proposed solutiOn 1S the use of a spht shaft des1gn separatmg the 

shaft mto two parts as shown m F1gure 5.2. The f1rst part (11) 1S the mam 

70 
--..J 

f-" 

!l~ 

f__ ~ 

@® 

~~@/~ --- f-JI 

Figure 5.1 

The STM4-120 Variable Swashplate Engine 

1) RotatIng shaft seal 

2) Thrust bearIng 

3) Stroke converter 

9) Front maIn bearIng 

10) Front crankcase castIng 

11) Cooler 

4) MaIn shaft 

5) Rear crankcase castIng 

6) Rear maIn bearIng 

7) Bevel gear 

12) Regenerator 

13) Heater 

14) Hot connectIng duct 

15) PIston 

8) Swash plate 16) CylInder 

17) Power control valve 

18) 011 scraper 

19) Cap seal 

20) RotatIng seal hOUSIng 

21) Pressure hull 

22) Cross-head 

23) Cross head bridge 

24) HydraulIc accumulator 
"'-I 

N 

j 111,11 

~ --

Figure 5.2 

The STM4-120 Variable Swashplate Stirling Engine with a Split-shaft Design 

1) Rotat1ng shaft seal assemhly 10) Rotary actudtor 19) IIpater 

2) Drive shaft 11) Ma1n shaft 20) lIot IOnnect1ng duct 

3) Rotating shdft seal 12) Bevel gear 21) P1ston assembly 

4) Mating r1l1~ 13) Swashplate 22) Lyl1nder/regenerator hous1ng 

5) Ra d 1 a 1 11 P seal 14) front ma1n bearing 23) Itydraul1c serV1ce assemhly 

6) Thrust bear1ng 15) 011 pump module 21) Power control valve 

7) Pres~ure hull 16) Front crankcase l5) 011 srrappr/capseal ass'y 

8) Rear ma1n hearing 17) Cooler 26) rrosshead 

q) Rear rlan~Lilse 1 n) HegenPra tor 27) IIccumulator 
shaft to WhICh the swashplate (13) IS attached. ThIS mam shaft IS supported by 

the mam bearmgs (8) and (14) and is subjected to the bendmg moment. The 

second part (2) IS the dnve shaft, subjected only to thrust load, which 

carnes the seal (3) and ItS matmg nng (4). 

The mam shaft (11) and the dnve shaft (2) are splmed together. The 

splme permIts the mam shaft to deflect freely WIthout transmIttmg transverse 

load or moment to the dnve shaft. The latter runs true WIth no deflectlOn at 

the seal. The splme clearance accomodates the slIght mIsalIgnment due to the 

deflectlOn of the mam shaft. 

In additlOn, JudiCiOUS selectlOn of the cross-sectlOnal dimenSlOns of the 

vanous sectIons of the mam shaft greatly reduces ItS deflectlOn at the front 

end where the bushmgs of the 011 pump module (15) are located. 

Under thIS program a vanable swashplate dnve deSIgn was completed whIch 

incorporates the splIt shaft and a separate end seal assembly. The deSIgn was 

based on a reference vanable swashplate deSIgn (STM4-120) prevlOusly generated 

by StIrlmg Thermal Motors, Inc. 

The deSIgn, set forth m the set of drawmgs attached to thiS report, 

achIeved the followmg results: 

(0 At the splIne the mam shaft deflectlOn IS 0.10 mm and the slope IS 

0.11 degrees. At the pump module end both deflectlOn and slope of the 

mam shaft vanish. 

(2) The splme clearance was chosen to accomodate the 0.1 mm deflectlOn 

and 0.11 degree slope of the mam shaft. CompressIve stress on the 

SIdes of the splme teeth determme ItS lIfe to be more than 10,000 

hours at full load. 

73 
(3) WIth the mam shaft allowed to deflect freely, the calculated mam 

bearmg's lIfe IS 13,000 hours at full load. 

The shaft deflectIOn analYSIS, bearmg lIfe and splme streneth 

calculatIOns are presented m the followmg sectIons. 

5.2 - Main Shaft DeflectIOn AnalYSIS 

It may be recalled that one of the desIgn requIrements IS that the 

deflectIOn and slope at the front end of the shaft vanish. The energy method IS 

used to evaluate the cross sectIonal dImenSIOn of the mam shaft whIch satIsfIes thIS requIrement. A free body dIagram of the main shaft and ItS model IS 

shown m FIgure 5.3. The mam shaft has five dIfferent sectIons as shown m the 

fIgure. SectIOn 114 was chosen for adJustmg the deflectIon slope at the front 

mam bearmg smce It IS the closest to that bearing, has the hIghest mfluence, and IS relatIvely free from other deSIgn constramts. 

The mam bearings are treated as SImple supports where the deflectIon 

vamshes and hence vamshmg slope at the front mam bearmg means vanIshmg 

deflectIOn and slope m the entIre regIon to ItS right. 

Coordmates and dImenSIOns are defmed m FIgure 5.3. Table 5.1 presents 

the POSItIons, cross sectIonal dImenSIOns, and sectIOnal moments of mertIa of 

the variOUS shaft model sectIons. 

A VIrtual moment (m) IS applIed at the front mam bear mg. The deflectIOn 

slope at that pomt (e) IS determmed by differentIatIOn of the elastIc stram 

energy WIth respect to the VIrtual moment and evaluatmg the denvative for a 

vanIshmg VIrtual moment (m = 0). Demandmg the slope to vanish YIelds one 

equatIon for the unknown moment of mertia of SectIon 4- (I 4) of the mam 

shaft. 

74 
a :: 111 .5 mm 

b :: 12R.5 mm 

c :: 145.5 mm 

d :: 164.8 mm 

t :: 17R.3 mm 

1-

----- I 

Rear ......... - ..... ---j.~ Front 

,------ - --- -- i._ _ _ _ _ _ _ _ _ _ _ __ _ 

x 

Section 1 2 3 

b 

c 

d 

Figure 5.3 

Free Body Diagram of Main Shaft 

M 

a 

4 

1 

5 

M /t 

a 

-------

Shaft model IS shown underneath. MaIn bearIngs are consIdered sImple supports. 
Table 5.1 

Shaft Model Cross SectIonal DImensIons 

Section Po sit ion ( mm ) OD(mm) ID(mm) Moment of Inertialmm41 

1 0< x< 111.5 - - 32.0 11.0 5.075 x 104 

2 111 • 5 ~ x = 128.5 38.0 11.0 1 .016 x 105 

3 128.5~ x~145.5 38.0 14.0 1 .005 x 105 

4 145. 5 ~ x ~ 164 .8 TBD 14.0 TBD 

5 164.8:: x~ 1'78.3 40.0 14.0 1 .238 x 10 5 

76 
The moment funct10n 1S: 

{

(M + m) xl£. 

M(x) = a 

(M + m) xl I.. - M 

a a 

The stram energy: 

2 

U = I I.. 1. M (x) dx 

o 2 EI 

@ 

@ 

The deflectIOn slope at the front mam bearmg: 

(5.1) 

(5.2) 

(5.3) 

SubstItutIOn of 5.1 m 5.3, carrymg out the dIfferentIatIOns, takmg mto 

account the dIfferent sectIonal moments of mertIa and settmg e = 0 YIeld the 

equatIons for I 4: 

M {1 a 2 1 b 2 2. - I (~) dx + - I (~) dx + 2- I

C 

~ (~ - 1) dx + 

E 11 0 1. 12 a 1. 13 b Jl. I.. 

1 d 1 Jl. } + - f ~ (-Jl.x - 1) dx + - I -Jl.x (-Jl.x - 1) dx ::: 0 

lu C fl. 15 d (5.4) 

The expreSSIOn m the brackets can be made to vanIsh mdependently of the 

magnItude of the applIed moment and the shaft materIal. SubstItutIon of the 

values for 11 ' I 2' I 3. and 15 from Table 5.1 and solvmg 5.4 for I 4 YIelds: 

= 6580 mm 

WIth an 10 of 14 mm, thIs corresponds to an 00 of 20.4 mm for sectIon 4. 

WIth the shaft geometry thus fully determmed the moment functIon IS 

integrated tWIce to Y1eld the deflectIon slope and the deflectIon. 

77 
The IntegratIon IS car ned out WIth the folloWIng quantItIes: 

ApplIed moment M = 1100 N.m (correspondIng to full load) a 

ElastIc modulus E = 2.0 x 10 (AISI 8620) 

The two requIred IntegratIon constants are supplIed by the two boundary 

condItIOns WhICh requIre that the deflectIOn at the two maIn bearIngs vanish. 

The results are shown In FIgure 5.4 and tabulated In Table 5.2. 

5.3 - BearIng Life CalculatIons 

The functIons of the maIn shaft bearIngs In the STM4-120 are to provIde 

radIal support for the maIn shaft whIle allOWIng angular deflectIons (slope) at 

the bearIng, and to provIde aXIal locatIOn of the shaft. The latter functIOn 

requIres a beanng WIth some thrust capacIty to prevent aXIal dIsplacements of 

the shaft. Sphencal roller bearIngs are an Ideal choIce for thIS applIcatIon 

sInce they have hIgh radIal load capacIty, some thrust load capacIty, and allow 

small angular deflectIons wIthout adversely affectIng performance. 

The mam shaft bearings specIfled are SKF 22208C spherical roller bearmgs 

wIth a BasIc DynamIc CapacIty (C 1O) of 63,600 N. 

From the free-body dIagram below, the bearIng reactIons are, 

= M h. a 

The maXImum value of the applIed moment In the STM4-120 engIne IS M = 

1100 Nm, and the bearing span, 9, = 178 mm, YIeldIng beanng reactIOns 

R 1 = -R2 = 6180N. 

The L 10 lIfe, m hours, for roller bearIngs IS: 

C 10/3 

16667 (~) 

N P 

78 
Moment,Mx 

Slope's!Y "dx 

0.102 mm 

Deflection; 

y 

'" ~ , 

I] ~~---- ~ ~ 

-c= - - - - - - - - - - ;:=..=.. 

~ \~ - - - - .-..l 

- -

----_ .... - --------=-= - - ------ ------- ------- - - --- - - - - - -

---------

V 

~ -- -- I --... I 

, \ I 

I 

I 

/ I 

i 

• 

, 

"" V V-

_/ V 

o III!! 129.5 145.5 178.-3 

Figure 5.4 

Mam Shaft Moment, Slope and DeflectIOn at Full Load 

79 

793 Nm 

0 

-3 07 Nm 

2.253 

Rad 

0 

-1.996 x 10-3 

P.adla ns 

0 

0.11 Omm 

~ 

X 
Table 5.2 

Main Shaft DeflectIon (y) and Slope (dy/dx) at Various PosItIOns (x) 

Along the AXIS at Full Load 

x y ~ dx 

(mm) (mm) (RAO) (OEG) 

0 0 -1.996xl0 -3 -0.1144 

10 -0.0199 -1.967xl0- 3 -0.0027 

30 -0.0391 -1.879 -0.1076 

40 -0.0736 -1.732 -0.0992 

50 -0.0875 -1.262 -0.0723 

60 -0.0989 -0.939 -0.0538 

70 -0.1061 -0.557 -0.0319 

75 -0.1084 -0.344 -0.0197 

82.45 -0.1097 0 0 

90 -0.1082 +0.382 +0.0219 

100 -0.1017 +0.940 +0.0539 

111 .5 -0.0868 +1.655 +0.0948 

115 -0.0808 + 1 .771 +0. , 015 

120 -0.0715 +1.943 +0.1113 

125 -0.0614 +2.123 ·0.121E 

128.5 -0.0537 T2.253 +0.1291 

130 -0.0504 +2.232 +0.1279 

135 -0.0394 +2.164 +0.1240 

140 -0.0287 +2.104 ·0.1205 

145.5 -0.0173 +2.046 +0.1172 

150 -0.0096 +1.423 +0.0815 

155 -0.0040 +0.839 +0.0480 

160 -0.0010 +0.367 +0.0211 

164.8 -0.0001 +0.022 +0.0012 

170 -0.0000 +0.008 +0.0005 

175 -0.0000 +0.001 +0.0001 

178.3 0 0 a 

80 
where 

N = speed m RPM 

C 10 = BasIc Dynamic CapacIty 

P = maXImum radIal load 

The maXImum speed of the STM4-120 IS 300 RPM, therefore, the L 10 lIfe of 

the mam bearings IS: 

16667 63600 N 10/3 

L10 = 3000 (6180 N ) 13200 hours 

The drive shaft In the STM4-120 transmIts the shaft torque through the 

wall of the pressurIzed crankcase. The crankcase pressure, contained by a 

mechanical face seal, exerts a thrust force on the dnve shaft equal to the 

crankcase pressure tImes the seal area. 

The specIfIed thrust bearing IS a TImken TTHD-Tl750 tapered roller bearmg 

wIth a BasIc Dynamic Thrust Rating (C90 ) of 32,200N at 500 RPM. 

The thrust load, T ,due to the crankcase pressure, IS: 

e 

T = P A = P (~d 2) ems m 4 s 

where 

Pm = crankcase mean pressure = 11 MPa 

A = seal area 

s 

d = seal dIameter = 37 mm 

s 

Therefore, the thrust load IS 

T e = (11 MPa) ( * ) (37 mm)2 = 1l,830N 

81 
The L 10 lIfe of tapered roller bearIngs can be calculated from the 

following equatIOn from the Tlmken Engineering Journal 

of: 

C90 x SF 10/3 

L 1 0 = ( T ) (3000 hrs) 

e 

where 

C90 = BasIc Dynamic Load Rating 

SF = speed factor = ( 500) 3/10 

rpm 

T = Load 

e 

= (~)3/IO 0584 3000 = • 

SubStItuting the approprIate values In thIS equatIon YIelds an L 10 lIfe 

L (32200 x .584 ) 10/3 (3000) = 14,060 hrs. 

10 11830 

5.4 - SplIne Strength CalculatIOns 

The torque capacIty of Involute splines can be calculated USing standard 

formulas and applIcatIOn factors, as In Machinery's Handbook, 22nd EdItIOn, pp 

922-927. The torque capacIty of a splIne IS determined by evaluating ItS 

strength wIth regard to three types of spline faIlure; shear faIlure under the 

roots of external teeth and/or at the pItch dIameter of the teeth, compressIve 

faIlure at the faces of the teeth, and burst faIlure of the Internal splIne. 

Internal splines may burst due to three kInds of tensIle stress; (I) tensIle 

stress due to the radIal component of the transmItted load; (2) centrIfugal 

tensIle stress; and (3) tensIle stress due to the tangentIal force at the pItch 

82 
line causing bending of the teeth. 

The splIne chosen to mate the main shaft wIth the dnve shaft has 30 

teeth, a 0.75 module and a 30° pressure angle. The tolerance class was chosen 

to allow suffIcIent clearance to accomodate the deflectiOn of the main shaft. 

The splIne IS desIgnated as follows: 

INT/EXT 30Z x 0.75 m x 30.0 R x 5H/5d 

ANSI B92.2M - 1980 

The stresses In the desIgnated spline are calculated below. 

Shear Stress Under the Roots of External Teeth: 

For a transmItted torque T, the torsiOnal shear stress (,rI

) Induced In the 

shaft under the root dIameter of an external splIne IS: 

T = 1 

where 

16 T Ka 

D 3K = 

re f 

5 16 x 1.25 x 10 N mm x 1.0 

x (20.81 mmP x 1.0 

T = transmItted torque = 125 Nm 

D

re = minor dIameter = 20.81 mm 

= 71 MPa 

Ka = spline applIcation factor for a uniform power source and 

Uniform load = 1.0 

Kf = spline fatIgue lIfe factor for 10,000 torque cycles = 

1.0. (a torque cycle consIsts of one start and one stop, 

not the number of revolutiOns.) 

Since the maximum allowable shear stress for case hardened steel (58-63 

Rc) as specifIed for the STM4-120 IS 345 MPa, the splIne strength wIth regard 

83 
to shear stress under roots of the external teeth IS more than suffICIent. 

Shear Stress at the PItch DIameter of Teeth: 

The shear stress at the pItch Ime of the teeth for a transmItted torque, 

TIS: 

1" 2 

4 x 1.2 x 105N mm x 1.0 x 1.0 = 

22.5 mm x 30 " 22.5 mm x 1,062 mm x 1.0 

31 MPa 

where 

T = transmItted torque 

K = splIne applIcatIOn factor 

a 

K = load dIstributIOn factor for mIsalIgnment of flexIble 

m 

splmes for mIsalIgnment of .002 m/mm and 22.5 mm face 

width = 1.0 

D = pItch dIameter = 22.5 mm 

N = number of teeth = 30 

L e = splme effective length = full length of splme for pitch 

dIameters less than 25 mm = 22.5 mm 

t = actual cIrcular tooth thIckness = 1.062 mm 

Kf = splIne fatIgue lIfe factor 

Agam, the mduced stress of 31 MPa IS much lower than the maXImum 

allowable stress of 345 MPa, hence the splme strength IS more than suffICIent 

WIth regard to shear stress at the pItch diameter. 

84 
CompreSSIve Stress on the SIdes of the Spline Teeth: 

Allowable compressIve stresses on splmes are very much lower than for 

gear teeth since non-umform load dIstributiOn and mIsalignment result m 

unequal load sharmg and end loadmg at the teeth. 

For a transmItted torque, T, the compressIve stress on the sIdes of the 

splme teeth IS: 

2TK K 

m a 

"c = D N L m K 

e w 

where 

5 2 )( 1.25 )( 10 N mm )( 1.0 )( 1.0 

= 22.5 mm )( 30 )( 22.5 mm )( 0.75 )( 0.7 31.3 MPa 

m = splme module = 0.75 (the module IS the inverse of the 

diametral pItch) 

Kw = wear life factor for flexIble splmes = 0.7 for 10,000 

9 hours at 3000 RPM (1 X 10 revolutiOns) 

Here the maXImum compreSSIve stress IS less than the allowable compressIve 

stress of 35 MPa, mdicatmg adequate performance for the deSIgn life of the 

splme. 

Burstmg Stress on the Internal Spline: 

For a transmitted torque, T, the radIal load tensIle stress IS: 

T tanq> 

ltDt L 

w 

= 1.25 )( 105 N mm )( tan 30.0° 

1[ )( 22.5 mm )( 9.0 mm )( 22.5 mm 

85 

5 MPa 
where 

, = pressure angle 30.0 

t = wall thIckness of mternal sphne = 1/2 (shaft OD - major dla.) = 9 mm 

W 

L = full length of sphne = 22.5 mm 

the centnfugal tenstle stress IS 

where 

DOl = OD of splme sleeve in mches = 1.563 

Dn = major dIameter of mternal splme m mches = #.~ = 1.563 

02 = 1.656 )( 10-6 )( 30002rpm)( (1.563 2 + 0.212 )( 0.9342) = 39 psi = 0.3 MPa 

and the beam loadmg tensIle stress IS 

4 )( 1.25 )( 105 N mm = 30 ~Pa 2 (22.5 mm) )( 22.5 mm )( 1.5 

where Y = LewIs form factor = 1.5 (for 30.0 pressure angle) 

Total tensIle stress tendmg to burst the rim of the external member: 

The total tenstle stress of 35 MPa IS much lower than the allowable 

tensIle stress of 380 MPa, mdlcatmg suffIcIent sphne strength wIth regard to 

burstmg. 

86 
APPENDIX I 

PISTON RING LEAKAGE EFFECTS 

ConsIder the upper pIston ring of the set of 2 pumping pIston rings shown 

In FIgure 2.1. 

The upper aXIal face of that ring IS exposed to a harmonically variable 

pressure (p ) wIth a mean value (p ) and an amplItude (Ap). u m 

The lower face of the ring IS exposed to a constant pressure equal to the 

maXImum variable pressure: 

Pu ::: Pm + t. P cos w t 

where W IS the angular speed of the rig and t 1S tIme. 

For laminar flow, leakage across the ring (defined as posItIve from the 

lower face to the upper face) IS proportIOnal to the pressure dIfference 

(p - p ) and Inversely proportIonal to the laminar resIstance (R). The 

e u 

leakage mass flow rate (m) IS thus gIven by: 

1 t.p ( ) rh = R (PI - p) = If 1 - co~ Ul t 

The leakage adds gas mass to the variable volume above the pIston. That 

added mass at any pOint (A m) IS found by tIme integratIon of the mass flow 

rate (m) gIven by EquatIOn 1.3. 

(I. 1) 

(1.2) 

0.3) 

t.m f rh dt t.p (wt - sin w t +<p) = wR (1.4) 

where <p IS an (yet undetermined) integratIon constant. 

87 
The mass mcrease (llm) raises the pressure p so that the disturbed u 

pressure (pu' ) m the volume above the nng is: 

pI = P (1 + ~m) u u mo 

where m is the undisturbed total gas mass m that volume. 

Substitution of (1.1) and (1.4) m (1.5) Yields: 

P~ = (Pm + ~p cos wt) [1 + m ~~R (wt - sin wt +q» ] 

o 

Since gP and ;t are small quantities higher order terms m these quantities 

m 0 

may be neglected with the followmg result: 

pI 

U 

Pm 

_ p + ~p [cos wt + (wt - sinwt +q»] m m wR 

o 

The rmg is constructed so that when p , IS larger than p the nng stops to 

u e 

(1.5) 

(1.6) 

0.7) 

seal and the extra mass m the variable volume will revert to the space below 

the nng. This happens just when wt = 0 and hence: 

~m = 0 

p' = p + ~p u m 

@ 

@ 

wt = 0 

wt = 0 

(a) 

(b) 

ApplicatIOn of (1.8) to (1.7) or to (1.4) determmes the mtegration constant: 

q> = 0 

and the disturbed pressure: 

Pm 

p' = p + ~p [coswt + (wt - sinwt)] 

u m m wR 

o 

88 

(I.8) 

(I. 9) 

(1.10 ) 
The pressure vanation m the volume above the rmg is caused by polytropic 

compreSSion. The volume (V) vanes harmOnically with a mean value (Vm) and an 

amplitude (f1 V) due to the motion of the piston: 

v = V - t:N cos w t 

m (1.11 ) 

The polytropic process enforces the followmg relatIOn between the volume 

(V) and the pressure (p ): u 

p Vn = p V n 

U m m 

where n is the polytropic constant. SubstitutIOn of (1.1) and (1.11) 

m (1.12) Yields: 

1 + l1p coswt = (1 

Pm 

l1V n 

cos wt)-

V 

m 

Since ~~ is a small quantity the nght hand side of (1.13) may be 

expanded m a senes With only thie first order term retamed: 

and thus: 

1 + l1 P cos wt 

Pm 

l1V 

n--

V 

m 

1 + n 

l1V - V cos wt 

m 

(I. 12) 

(1.13 ) 

(1.14) 

(1.15) 

The work (W Lp) performed (per cycle) by the gas undergomg the polytropic 

compreSSIOn distrubed by leakage is given by mtegratIOn over the cycle of: 

dV 

= P~ d(wt) d(wt) (1.16) 

89 
Conveniently defInIng a. _ wt and IntegratIon (1.16): 

WL P = f2'1{p 

0 m + 6p [COl> a + ~ m w R (a - slna)J}IN l>lna do .. - 3'16p 6V ~ m wR o 0 

(1.17) 

The equatIon of state for the (Ideal) gas requires that: 

T 

( c - c ) --.!:!! 

p v V 

m 

(I.iS) 

where c and c are the specIfIc heat values of the gas at constant pressure and 

p v 

volume respectIvely and T IS the mean volume temperature. DefInIng y :: c Ic 

m p v 

and SubstItuting (1.15) In (I.17): 

Pm 'V _ 1 AV 2 = - 3. - n' C

p 

T <_U_) wR 'Y m V m (1.19) 

The same consIderation applIes to the lower rIng of the pumpIng rIng set and the 

vanable volume below It so that the total work performed by the working gas In 

the ng due to thIS leakage effect is 2W Lp. 

The total heat (QLp) dIsSIpated by the ng (per cycle) IS found by applYIng the 

energy equatIon (for a CyclIC process). 

(1.20) 

WhICh YIelds: 

Pm 'Y- 1 < 6V)2 6lt - n -- C T wR 'Y p m V 

m 

(1.21 ) 

An addItIonal loss associated wIth the pIston rIng IS the VISCOUS dISSIpatIOn In the 

leakage path. The volumetnc flow (4J) undergoes a pressure drop (pt - pu) and thus 

g() 
dISSIpates energy (W Lo) as folJows: 

. dWLo = ~ (Pi - pu) dt = ; ~p (t - cosa) da (1.22) 

The volumetrIc flow IS: 

(1.23) 

where P IS the mean gas densIty whIch, accordIng to the equatIon of state IS: 

m 

Thus: 

dWLo 

'Y Pm 

'Y- 1 C T P m 

'Y - 1 meT ~ p (1 _ cos a) da 

'Y wpm Pm 

SubstitutIOn of (1.3) and (1.15) In (1.25) and IntegratIOn over the cycle 

Yields: 

P 2 1 V 2 21 Pm 2 1 V 2 

W =...!!!. n !....:....- C T (~) f (1 - cosa)2 da = 31 - n l=-.. C T (~) Lo III R 'Y P m V 0 IIlR 'Y P m V 

m m 

DoublIng this amount to account for the other rmg of the set Yields the total 

heat (QLo) diSSipated (per cycle) m the rIg due to the resIstance of the flow 

path: 

The total leakage loss (Q L) per cycle is the sum of the above two 

91 

(1.24) 

(1.25) 

(1.26) 

(1.27) 
effects: 

P 2 

::; QLp + QLo ::; 6. ~ n (n + 1) 'Y - 1 C T (~ v) QL w R 'Y P m V 

m 

(I.28) 

The "leakage coefficient" (CL ) defined m Section 2.7.2 above should be given by: 

Pm 2 

::; 6lt - n (n + 1) 'Y - 1 C T ( ~ V) 

R 'Y P m V 

m 

The ratio of the orifice effect to the phase shift effect IS found by dlvldmg 

(1.27) by (1.21) to be equal to the polytropIc constant n: 

92 

(1.29) 

(1.30) 
APPENDIX II 

PISTON RING FRICTION 

ConsIder one of the two pumping pIston rings (say the upper one) of a UnidIrectIOnal 

set in a double acting ng and the vanable volume above It. The ring IS moving Up and 

down WIth a substantIally harmOniC motIon of stroke (S) so that ItS instantaneous 

POSI tIOn (x) IS: 

S 

x = "'2 cos w t (11.1) 

where W IS the angular speed and t IS tIme. 

ThIS motIon of the pIston changes the volume (V) above the ring as follows: 

V = V - I:::. V cos w t 

m 

where V IS the mean volume and I:::. V IS the volume vanatIOn amplitude. m 

(11.2) 

The gas IS being compressed in a polytropIc process WIth constant (n) so that ItS 

instantaneous pressure (p) IS: 

p = p + I:::.p cos w t 

m 

where p IS the mean pressure and I:::.p IS the pressure vanatIOn amplitude m 

The relatIOn between the instantaneous volume (V) and the instantaneous 

pressure (p) for a polytropIc process WIth constant (n) IS: 

pVn 

= V n const. = p mm 

SubstItutIOn of (2) and (3) in (4) yields: 

or: 

P V n (1 + ~ cos wt) (1 I:::. V cos w t)n = p V n 

m m Pm - Vm m m 

1 + ~ cos wt = 

Pm 

I:::. V n (1 - V cos wtf 

m 

93 

(11.3) 

(11.4) 

(11.5) 
flV The right hand sIde may be expanded mto a series wIth hIgher order terms m ( V) 

neglected: 

I + ~ cos wt 

Pm 

flV = I + n V m cos wt + H.O.T. 

ThIS YIelds the followmg relatlOn between the relatIve pressure and volume 

VarIatlOn amphtudes: 

m 

(11.6) 

The pressure on the other sIde of the rmg and m the groove IS the maXImum 

pressure of the variable volume: p + flp. m 

The outward normal force (N) on the rmg, forcmg It at the wall IS (as seen m 

FIgure II-1): 

N = TI Dh[(p + flp) - t {(p + flp) + (p + flp cos w t)}] = t TI Dhflp (I - cos w t) (11.7) m m m 

where D IS the mean dIameter of the pIston rmg and h IS ItS aXIal thIckness. 

The frIctlOn force (F) IS proportIonal to the normal force (N) wIth the frIctlOn 

coeffIcIent (f) as the proportlOnalIty constant. F acts always m a directlOn 

OpposIte the pIston speed (x) 

F = t TI Dhflpf (I - cos w t) 1 sm w t 1 

sm wt 

The work (W) performed agamst the frictIon IS: 

(II.8) 

dW = -F dd~ d a = - tTIDhflpf(I - cosa) I::~ ;1 (- ~ sma) da = 

= t TIDhSfllp(I - cosa) 1 sm aida (II.9) 

where a _ wt 

94 
p = p + Llp cos w t 

m 

p = p + Llp m 

F = 1I'0hp = 11' Oh (p + tlp) 1 m 

F = t Dh(p + p) = 11' Oh [p + tLlp (1 + cos w t)] o m 

N = F - F = t 11' OhLlp (1 - COS w t) 1 0 

I r---+ D 

Figure D.I 

Lateral Gas Forces on a Pumping Piston Ring 

p = Instantaneous vanable gas pressure 

p = Maxlmum pressure 

F = Pressure force acting on the 00 of the ring o 

F = Pressure force acting on the 10 of the ring 1 

9S 

F. 

1 

-
The work per cycle IS achIeved by integratIon of (9) over the entIre cycle: 

2lT 

W = J dW da d a = o 

lT dW 2 .r dad a = IT DhSf~p 

o 

(II. 10) 

The same consIderatIOn applIes to the other ring and hence the total pIston 

ring frictIon loss (Qf) In the rig IS tWIce the amount of equatIon (lO). Also, 

SubstitutIon of (6) YIelds: 

Q 27T DhSfp n ( ~ V ) f = m V 

m 

96 

(II.li) 
Construction 

APPENDIX III 

Metallographic Inspection 

of 

Heat Pipe #2 

Coulson M. Scheuermann 

NASA Lewis Research Center 

Cleveland, Ohio 

This heat pipe contained no getter; therefore, no getter bucket or tube was included 

in its construction. The 27 112 inch (70 em.) long tube body of the heat pipe was of type 

321 stainless steel tube stock, 1 112 inch (3.8 em.) diameter with a 0.049 inch (0.12 em.) 

wall thickness. 

Top and bottom caps were machined from type 304 stamless steel plate stock. The 

bottom cap was TIG-we1ded to the tube body. The top cap had a 114 inch (0.64 em.) 

diameter type 304 stainless steel evacuation tube welded into it. This small tube had a 

valve attached to allow air addition. 

The wick was a rolled tube of type 316 stainless steel, 0.002 inch (0.005 em.) 

diameter wire, 165 mesh screen. The wick was inserted into the bottom of the tube body 

and held in place by a coiled 1116 inch (0.16 em.) diameter nichrome wire spring. 

Sixteen grams of sodium were dropped into the pipe prior to closure. The top cap 

was then TIG-we1ded in place. 

Operation 

Heat pipe #2 (figures 4.1 and 4.2, also table 4.0 operated for a total 3,142 hours at 

800°C. Starting at 2,158 hours of operation, a series of doses of air, about 2.0 Ncc each, 

were bled into the heat pipe. About 116 Ncc of total air volume was introduced. Full 

heat pipe operation recovered within 15 to 20 hou'-s after each dose. 

97 
Metallographic Inspection 

The heat pipe was cut mto sections for visual inspection as shown m figure III-I. 

Specific areas were then chosen for chemical and metallographic analysis. 

The dark deposit visible near the bottom of the heat pipe in figure III-I, and 

enlarged in figure III-2, was removed for chemical analysis. 

SectlOns from the heat pipe were cut for metallographic inspection from the 

condenser and various locations down to the bottom end cap. These pieces were given a 

coating of copper, mounted in epoxy and exammed using an electron microscope. 

Results and Discussion 

Chemical analysis of the dark deposit in figure III-2 revealed it to be primarily 

copper, with minor amount of iron, manganese, chronium and sodium. It is believed that 

this was caused by the inadvertent introduction of copper filmgs from tools used during 

the pipe constructlOn. 

Metallographic and electron microscopic examination of samples from the heat pipe 

condenser end did not reveal any serious structural problems. There was a slight surface 

effect extending to a depth of about 0.0009 inch (0.002 cm.), especially along grain 

boundaries (figure III-3). This effect was general along the entire heat pipe wall. It 

seemed related to a loss of Ni, and possibly, C from the wall. 

More extensive attack by the sodium appeared in the evaporator section at the 

bottom of the pipe. The nichrome wire in this location suffered grain boundary attack and 

removal of Ni and C to about 0.002 inch (0.006 cm.) depth in the region where the wire 

contacted the wick, but none away from the Wlck. The wick itself was highly attacked 

(figure III-4). The bottom end cap experienced a greater degree of attack than did the 

pipe wall. Here the attack was to a depth of about 0.002 inch (0.006 cm.) depth. 

98 
Considering the intentional lack of control over cleanliness and sodium purity, the 

operation of this heat pipe demonstrated the ruggedness and tolerance of this system. 

Even though there were obvious plugged and corroded portions of the wick in the 

evaporator area, the pipe did not exhibit signs of operational distress. Corrosion, although 

present, did not hamper the operation, nor did it proceed to the extent where it could 

cause failure. Even intentional air leakage did not seriously hamper operation, since the 

pipe operation soon recovered to pre-leakage performance. 

This pipe had many different materials in its construction. If all parts were made 

from a common, relatively corrosion resistant material, potential problems resulting from 

corrosion could be minimized. 

Conclusions 

The main conclusions to be reached from these test results are: 

1. Corrosion did occur, but it was not severe enough to cause failure. It did not appear 

to be a serious problem for the pipe wall or spring; 

2. Corrosion of the wick in the evaporator section could become an operational 

problem; 

3. Potential corrosion problems could be reduced by using a single material of 

construction for all heat pipe parts; 

4. Satisfactory heat pipe operation can be achieved without meticulous cleanliness or 

sodium purity control. 

99 
Figure 111-1. 

Overall view of sectional heat pipe H2. 

Figure 1II-2. 

View of sectioned heat pipe H2 

showing the dark deposit. 

100 
F i g.u reI 1 1 - 3 . 

A typical electron photomicograph of the 

wall of heat pipe 12 in the condenser section 

showing the slight corrosion attack. 

I . 

I~~ 

Figure 111-4. 

An electron photomicrograph of the corroded 

wick in the evaporator section of 

heat pipe 12. 

101 
1 Report No 2 Government Accession No 

NASA CR-174994 

4 Title and Subtitle 

Experimental Assessment of Advanced Stirling Component Concepts 

7 Author(s) 

Benjamin Ziph 

9 Performing Organization Name and Address 

Stirling Thermal Motors, Inc. 

2841 Boardwalk 

Ann Arbor, Michigan 48104 

12 Sponsoring Agency Name and Address 

U.S. Department of Energy Office of Vehicle and Engine R&D 

Washington, D.C. 20545 

15 Supplementary Notes 

3 Recipient's Catalog No 

5 Report Date 

October 1985 

6 Performing Organization Code 

8 Performing Organization Report No 

10 Work Unit No 

11 Contract or Grant No 

DEN 3-351 

13 Type of Report and Period Covered 

Contractor Report 

14 Sponsoring Agency Coda Report No. 

DOEINASA/0351-1 

Final Report. Prepared under Interagency Agreement DE-AIOl-77CS51040. Project 

Manager, J. Ca1rell1, Power Technology Division, NASA Lewis Research Center, 

Cleveland, Ohio 44135. 

16 Abstract 

This final report presents the results of an experimental assessment of some 

advanced Stirling engine component concepts. Under this program, high performance piston rings, reciprocating oil scrapers and heat pipes with getters and 

with mechanical couplings were tested. The tests yielded the following results: 

(1) Bonded, split, pumping piston rings, in preliminary testing, proved a promising concept, exhibiting low leakage and friction losses. Solid piston 

rings proved impractical in view of their sensitivity to the operating 

temperature. 

(2) A babb1t oil scraper in a compliant housing performed well in atmospheric 

endurance testing. In pressurized tests the scraper did not perform well 

as a containment seal. The latter tests suggest modifications which may 

adapt T1 successfully to that application. 

(3) Heat pipe endurance tests indicated the adequacy of simple, inexpensive 

fabrication and filling procedures. Getters were proved to increase the 

tolerance of the heat pipes to the presence of a1r,and commercially available couplings were demonstrated to be suitable for heat pipe application. 

In addition to the above tests, the program also included a design effort for a 

split shaft applicable to a swashplate driven engine with a pressurized crankcase. The design is aimed, and does accomplish, an in~rease in component life 

to more than 10 000 hr. 

17 Key Words (Suggested by Author(s» 

Stirling engines; Heat pipes; Piston 

rings; Scraper seals; Test data; Split 

shaft design; Swashplate 

18 Distribution Statement 

Unclassified - unlimited 

STAR Category 85 

DOE Category UC-96 

19 Security Classlf (of this report) 

Unclassified 

20 Security Classlf (of this page) 

Unclassified 

21 No of pages 

105 

'For sale by the National Techmcallnformatlon Service, Spnngfleld, Virginia 22161 

22 Price' 

A06 
End of Document